High density combined cycle power plant process

ABSTRACT

A process for increasing the specific output of a combined cycle power plant and providing flexibility in the power plant rating, both without a commensurate increase in the plant heat rate, is disclosed. The present invention demonstrates that the process of upgrading thermal efficiencies of combined cycles can often be accomplished through the strategic use of additional fuel and/or heat input. In particular, gas turbines that exhaust into HRSGs, can be supplementally fired to obtain much higher steam turbine outputs and greater overall plant ratings, but without a penalty on efficiency. This method by and large defines a high efficiency combined cycle power plant that is predominantly a Rankine (bottoming) cycle. Exemplary embodiments of the present invention include a load driven by a topping cycle engine, powered by a topping cycle fluid which exhausts into a heat recovery device.

RELATED APPLICATIONS

This is a divisional of application Ser. No. 09/783,693, filed Feb. 14,2001 now U.S. Pat. No. 6,494,045, which was a divisional of applicationSer. No. 09/359,813, filed Jul. 23, 1999 now U.S. Pat. No. 6,230,480,which claimed the benefit of U.S. Provisional Application Nos.60/098,468, filed Aug. 31, 1998, and 60/125,576, filed Mar. 23, 1999.

FIELD OF THE INVENTION

This invention relates generally to combined cycle power plants that mayor may not incorporate cogeneration into their cycle. As will bedemonstrated by the following disclosure, the increasing need for moreenergy efficient and environmentally friendly methods of generatingpower has prompted a widespread search for systems and methods toachieve these goals. However, current technologies have a generallymyopic view of the total economic impact imposed by a concentration onenergy efficiency and environmental issues alone.

The present invention proposes to break with tradition and include aspart of the economic and environmental analysis the complete equipmentcomplement required to implement a desired plant load (power) rating. Byincorporating this analysis into a new system and method of supplementalfiring and heat recovery, the present invention dramatically cuts theoverall economic and environmental cost of installed power plants byreducing the equipment complement while maintaining or reducing plantemissions. The result of this improvement over the art is cheaper andcleaner electrical energy than would be possible using conventionalcombined cycle plants that are currently known in the art.

BACKGROUND OF THE INVENTION

Overview

Combined cycle power plants and cogeneration facilities utilize gasturbines (GT(s)) as prime movers to generate power. These GT enginesoperate on the Brayton Cycle thermodynamic principle and typically havehigh exhaust flows and relatively high exhaust temperatures. Theseexhaust gases, when directed into a heat recovery boiler (typicallyreferred to as a heat recovery steam generator (HRSG)), produce steamthat can be used to generate more power and/or provide process steamrequirements. For additional power production the steam can be directedto a steam turbine (ST) that utilizes the steam to produce additionalpower. In this manner, the GT produces work via the Brayton Cycle, andthe ST produces power via the Rankine Cycle. Thus, the name “combinedcycle” is derived. In this arrangement, the GT Brayton Cycle is alsoreferred to as the “topping cycle” and the ST Rankine Cycle is referredto as the “bottoming cycle,” as the topping cycle produces the energyneeded for the bottoming cycle to operate. Thus, the functionality ofthese cycles is linked in the prior art.

Rankine Cycle

Steam has been used for power applications for more than a century.Early applications utilized a pump to bring the water up to the desiredpressure, a boiler to heat the water until it turned to steam, and asteam engine, typically a piston type engine, to produce shafthorsepower. These power plants were used in factories, on locomotives,onboard steamships, and other power applications.

As technology progressed, the trend for the use of steam enginesdiminished and the use of steam turbines increased. One advantage of thesteam turbine was its overall cycle efficiency when used in conjunctionwith a condenser. This allowed the steam to expand significantly beyondnormal atmospheric pressure down to pressures that were only slightlyabove an absolute vacuum (0.5 to 2 pounds per square inch absolute(psia)). This allowed the steam to expand further than in an atmosphericexhaust configuration, extracting more energy from a given mass ofsteam, thus producing more power and increasing overall steam cycleefficiency. This overall steam cycle, from a thermodynamic perspective,is referred to as the Rankine Cycle.

FIG. 1 illustrates the thermodynamic operation of the Rankine Cycle. InFIG. 1, graph (100) illustrates the Rankine Cycle on a Pressure versusVolume plot. From point (101) to point (102), water is pressurized atconstant volume. From point (102) to point (103), the water is boiledinto steam at constant pressure. Point (103) to point (104) defines theprocess where the steam expands isentropically and produces work. Then,from point (104) to point (101) the low-pressure steam is condensed backto water and the cycle is complete.

Also in FIG. 1, graph (110) illustrates the Rankine Cycle on aTemperature versus Entropy plot. From point (111) to point (112), wateris pressurized. From point (112), the water is boiled into steam atconstant temperature until it is all steam, then it is superheated topoint (113). Point (113) to point (114) defines the process where thesteam expands isentropically and produces work. From point (114) topoint (111) the low-pressure steam is condensed back to water atconstant temperature to complete the cycle. See Eugene A. Avallone andTheodore Baumeister III, MARKS' STANDARD HANDBOOK FOR MECHANICALENGINEERS (NINTH EDITION) (ISBN 0-07-004127-X, 1987) in Section 4-20 formore discussion on the Rankine Cycle.

Power Plant Cycle

For a number of decades, the Rankine Cycle has been used to produce mostof the electricity in the United States, as well as in a number of othercountries. FIG. 2 illustrates a schematic of the basic Rankine Cycle,with the four primary components being the Boiler Feed Pump (BFP) (201),Boiler evaporator/superheater (BOIL) (203, 205), Steam Turbine (ST)(207), and the Condenser (COND) (209). Note that either one or multiplesof any component are possible in the arrangement, but for simplicity,only one of each is shown in FIG. 2. The sub-critical Rankine Cycle(steam pressures less than 3206.2 psia) starts as water at the inlet(211) of the BFP (201). The water is then pumped to a desired dischargepressure by the BFP (201). This pressurized water (202) is then sent tothe evaporator (EVAP) (203) where heat is added to the pressurizedwater. Typically this is accomplished by burning a fuel in the boiler,and the heat of combustion is then transferred to the pressurized waterthat is routed through tubes and other passages and/or vessels in theboiler. As sufficient heat is added to the pressurized water, it boilsand turns into steam (204). This steam now exists in the two-phaseregion where both steam and water coexist at the same pressure andtemperature, called the saturation pressure and saturation temperature.For most applications designed in recent decades, this steam (204) isthen sent to a superheater section (SHT) (205) in the boiler where it isheated to a higher temperature than saturation temperature. This steam(206) is now referred to as superheated steam. Superheated steam reduces(but does not eliminate) the risk of water carryover into the steamturbine (207), which is of concern since water carryover can causeextensive internal steam turbine damage. Of more importance, however, isthe fact that superheated steam yields better cycle efficiencies. Thisis of great importance to large central power stations.

Once produced, the superheated steam (206) is sent to the steam turbine(207), typically via one or more pipes. The steam then begins to expandin the steam turbine (ST) and produce shaft horsepower. After travelingthrough the steam turbine down to a low exhaust pressure, the steamexits the ST (208), and is sent to the condenser (209), where it is thencondensed back into water. This device is typically a tubed heatexchanger, but can also be other types of heat exchangers such as aspray chamber, air-cooled condenser, or other heat exchange device usedfor a similar purpose. After rejecting heat from the low-pressure steamand condensing the steam back to water, the condenser collects the waterin an area commonly referred to as the hotwell (HW) (210), where it isthen typically pumped through the condensate line (211) and back to theBFP (201). Shaft horsepower produced in the ST is converted intoelectrical power in the generator (GEN) (212). This cycle of one unit ofwater from the point of beginning, through the system, and back to thepoint of origin defines the basic Rankine Cycle.

Current power plants using only steam as the motive fluid typically usea boiler to produce the steam. This boiler may be fueled by a variety offuels, including oil, natural gas, coal, biomass, as well as others,such as nuclear fuel. The boilers may also use a combination of fuels aswell. Depending upon capital cost considerations, fuel costs,maintenance issues, and other factors, the owners and engineers willselect the steam pressure and temperature at which the boiler willproduce steam.

Due to the size and weight of large steam turbines, they requireextended periods for start-up. This is due to the thick metal casingsand large heavy rotors that are utilized in their construction.Therefore, these machines require long start-up periods to allow theseheavy components to warm up uniformly, and avoid interference betweenstationary and rotating parts that may occur due to differential thermalexpansion.

Although the heavy construction is a deterrent to rapid startup, itprovides for robust construction and sustained performance levels. Evenafter four (4) years of nearly continuous service, the performance decayfor a large ST should be less than 2%. This performance decay, combinedwith the fact that the boiler feed pumps only consume about 2% of the SToutput, mean that the performance levels for a ST sustain near optimumlevels for extended periods of time, even with decay in the auxiliaryloads (BFP). In other words, if the BFP efficiency decays from 75% to65%, the auxiliary load only increases from 2.00% to 2.31%. This is asmall effect on the net output of the Rankine cycle plant, and isanother one of its major advantages.

Brayton Cycle

The Brayton Cycle varies quite differently from the Rankine Cycle, as amajor part of the cycle involves the compression of the working fluid,which is a compressible gas. This process consumes a great deal ofpower, therefore, efficient compression of the working fluid isessential to an efficient Brayton Cycle.

Common engines that utilize a Brayton Cycle are aircraft turboprops, jetengines, and gas turbines for stationary application. These engines workby ingesting air (the working fluid), compressing it to a higherpressure, typically 3 to 30 times that of the surrounding ambient air,adding heat through direct combustion (although heat addition from anexternal source is also possible), and then expanding the resultinghigh-pressure hot gases through a turbine section. Aircraft enginesprimarily produce thrust to propel an aircraft through the air.Therefore, some or perhaps none of their output is in the form of shafthorsepower (a turboprop gas turbine engine may drive the propeller, butmay also produce some thrust from the high velocity exhaust gases).

For stationary gas turbine applications, the purpose of the engine is toproduce shaft horsepower. Approximately ⅔ of the energy produced by theturbine section of the gas turbine is required to drive the compressorsection, with the remaining ⅓ available to drive a load. This drawbackof GT systems may be used to advantage in the present invention asdescribed later in this document.

Aircraft engines utilize the Brayton Cycle because these engines offerhigh thrust-to-weight ratios. This is needed to minimize the aircraftweight so it can fly. For stationary applications, gas turbines are usedto provide electrical power at peak loads. This is another advantage theBrayton Cycle engines have over Rankine Cycle engines: rapid start andstop times (relatively speaking). Since steam turbines are large heavyengines, it is necessary to start them slowly, and allow the heat toslowly soak into the thick casings so as to avoid thermal distortion andpotential rubs between the stationary components and rotating componentsof the engine. A large power plant steam turbine may require a 24-hourwarm-up sequence from cold start to reach full load. However, due to thelower operating pressures and lighter weights, gas turbines can bestarted and brought to fall load within a matter of minutes of start-up.

Therefore, many utilities in the United States and other countries usegas turbines to provide electrical power during peak demand. Theseturbines are not very efficient in simple cycle (25% to 30% LHV), butmeet the electrical demand requirements for a few hours each day.

Steam Turbine Design

When designing a steam turbine for a power plant application (constantspeed), the steam turbine design engineer first examines the outputrating desired by the customer. This is because the steam turbine willbe custom designed and manufactured for the customer to hisspecification. The steam turbine will not be totally designed from aclean sheet of paper as may be inferred by “custom”, but will utilizecomponents from a “family” of hardware and have a unique steam path forthe application. After turbine rating, the ST design engineer will lookat the plant steam conditions, and based upon these parameters determinean inlet flow to the turbine high-pressure (HP) section. Utilizing thisinformation, the ST design engineer can select the optimum HP casing forthe application. In a similar fashion, he can also select the optimumintermediate pressure (IP) and low-pressure (LP) casings as well.

Knowing which casings to use, the engineer then selects the appropriateblading (both stationary and rotating) for the application. This bladingsize is determined primarily by the volume flow (as opposed to massflow) of steam through the turbine. With casings and blading determined,the engineer completes the ST design by selecting valves, controls,instrumentation, and other accessories required for operation of the ST.The final design is a high efficiency ST optimized for the customer'ssteam conditions and desired rating.

An interesting note concerning this design philosophy is that two STswith the same steam conditions but with large differences in rating (forexample, 200 MW versus 400 MW) may actually appear almost identical whenviewed from the outside. This is because the optimum casings selectedwere designed to cover the flow range of both units. However, due to thelarge volume flow differences, the large unit would have blades that areapproximately twice the size (height) internally. It is interesting tonote, however, that both these units might have nearly the same HP andIP casings. This means that the larger ST, even with a dramatic increasein rating, may be only incrementally more expensive to manufacture thanthe ST with the lower rating. This fact may be used to advantage in thepresent invention as described later in this document.

Gas Turbine Design

Unlike the steam turbine, the gas turbine is not a custom designedmachine for each customer. Although accessories such as the startingmeans, lube oil cooler type, and control options may be specified by thecustomer for a particular application, the core engine is essentiallystandard. Much of this is due to the fact that the gas turbine isactually a packaged power plant, which needs essentially only fuel toproduce power. In contrast, the steam turbine is merely a component of apower plant, and requires a boiler, BFP, and condenser to become acomplete power plant. Therefore, the gas turbine compressor section,combustion system, and turbine section must all be designed to worktogether. Since the design of the GT is a highly intensive engineeringtask, GT designs are generally completed and extensively tested, afterwhich they are mass produced without variation to the core enginedesign. This eliminates the customer's ability to specify power outputfor either a facility with gas turbines only or a combined cyclefacility in the prior art. When building a combined cycle plant, thecustomer simply must choose from a selection of standard offerings by amanufacturer that best meets his needs for power output, efficiency, andcost.

Steam Turbine/Gas Turbine Efficiency and Rating Comparison

The largest and most efficient GT available today for 60-cycle powerproduction is rated at approximately 250 MW with an efficiency of 40.0%LHV (Lower Heating Value). An example of this GT is the Westinghousemodel 501G. This is in contrast to STs that can be rated up to as highas 1500 MW and have overall cycle efficiencies in excess of 45% LHV.Therefore, comparing a Rankine Cycle power plant to a Brayton cyclepower plant, where each employs the largest and most efficient turbineavailable, the single ST Rankine cycle is approximately six (6) timeslarger in rating and 12.5% more efficient than the Brayton Cycle withits best GT. This fact may be used to advantage in the present inventionas described later in this document.

Cogeneration/Combined Cycle

One characteristic of the gas turbine is that it expels high volumes ofexhaust gases at high temperature. With the advent of the Arab oilembargo of 1973 and higher energy prices, more focus was put on findingways to utilize the energy contained in these high temperature exhaustgases.

Significantly higher energy prices in the early 1970s signaled the startof a wave of small power plants built using the principles ofcogeneration. Cogeneration is defined as the simultaneous production ofmechanical or electrical energy in conjunction with thermal energy. Inother words, the utilization of an engine (gas turbine or otherwise) toproduce power, while at the same time using waste heat from the enginefor another process, thus displacing fuel that would otherwise be usedfor said process. This was a very efficient method from a fuelutilization perspective and was encouraged by the United States PublicUtilities Regulation and Policies Act (PURPA) of 1978, which mandatedthat the local utilities must purchase power from qualifiedcogenerators, and buy it at a rate which included avoided cost for newpower plants.

At first cogeneration projects were small, typically less than 50 MW.They consisted of small gas turbines with a HRSG to produce steam. Inmany instances, the steam pressures were relatively low (less than 600psig), as the steam was used for process requirements. Some projectsincluded a steam turbine, while others did not. As the industry matured,larger plants with higher steam pressures were designed to increasebottoming cycle efficiency. In addition, the major gas turbinemanufacturers designed and built larger and more efficient gas turbinesto meet the needs of the cogeneration marketplace. Soon, due to theirhigh efficiency, low emissions, and low capital cost (dollars per kW ofcapacity), cogeneration power plants gave way to combined cycle powerplants (plants that produced only power and provided no useful thermalenergy as was the case with cogeneration plants). Some cogenerationprojects are still being proposed and constructed, but they are nowtypically referred to as combined heat and power (CHP) projects.

Although there was this gradual shift from small cogeneration projectsto large combined cycle power plants, the arrangement and overall systemand method for producing power was for the most part unchanged. The gasturbine(s) was the primary engine, and a HRSG was utilized to capturethe heat in the GT exhaust gases. Optimized for maximum powerproduction, the steam turbine(s) produced additional power equal toapproximately 50% of the power produced by the gas turbine(s). The HRSGwas typically a two or three pressure level boiler to maximize heatrecovery and steam turbine was designed to accept steam from allpressure levels of the HRSG. A review of the manufacturers standardcombined cycle offerings will illustrate this trend. The 1997TURBOMACHINERY HANDBOOK, (USPS 871-500, ISSN 0149-4147), tabulatesstandard combined cycle power plants available from variousmanufacturer's including ABB, General Electric, and Westinghouse. Inmost every instance, the steam turbine's output is within the range of40% to 60% of the gas turbine(s) output. General Electric informativedocument GER-3567G, 1996, “GE Heavy-Duty Gas Turbine PerformanceCharacteristics,” by Frank J. Brooks provides the output for the gasturbines used in their combined cycle power plants.

In summary, the system and method utilized by the major manufacturer'sof combined cycle power plant turbomachinery evolved from the smallcogeneration power facilities that were designed to produce both powerand thermal energy simultaneously. The sizes for combined cycle powerplants have grown from small cogeneration projects under 50 MW to largestructured plants producing in excess of 700 MW (as in the Westinghouse2X1 501G combined cycle). These plants are primarily gas turbine powerplants, with the steam turbine producing additional power which isnominally 40% to 60% of the power produced by its associated gasturbine(s). With the gas turbine as the prime engine, the ratings on thestandard combined cycle power plants are very rigid, as gas turbines areproduction line items, versus steam turbines which are largely customdesigned and manufactured. A new system and method that offers moreflexibility, without compromising the benefits of combined cycle powersuch as high efficiency, low emissions, and low capital cost, would bewelcomed by the industry.

DESCRIPTION OF THE PRIOR ART

Efficiency Optimizations

Feedwater Heater

With Rankine Cycle plants producing billions of dollars of electricityannually, and consuming commensurate amounts of fuel each year, a greatdeal of design and analysis has been done to optimize the Rankine Cycleby introducing small variations or revised configurations. FIG. 3illustrates some of the common variations that are used to design aRankine Cycle for optimum efficiency. Part (303) of FIG. 3 schematicallyrepresents a feedwater heater (FWH). This device is typically a shelland tube heat exchanger, but could be a plate and frame heat exchanger,vortex mixing heat exchanger that mixes the feedwater with small amountsof steam, or other heat exchange device used for a similar purpose.Analysis has proven that utilizing extraction steam from the steamturbine to preheat water before it enters the boiler increases the cycleefficiency.

The feedwater heater (303) uses steam that is extracted from the steamturbine at an optimum point to preheat the water between the condenser(319) outlet and the boiler inlet (306). A second feedwater heater (305)is shown in this example. The number of feedwater heaters and theiroptimum steam conditions are dependent upon a number of factorsincluding but not limited to steam turbine inlet pressure, steam turbineinlet temperature, reheat steam conditions, feedwater heatereffectiveness, and other factors. Typically, the number of feedwaterheaters, their design, and the inlet steam conditions for thesefeedwater heaters must be determined for each power plant due tovariations in each power plant's design and individual conditions.

Reheat

Another variation on the Rankine Cycle used to improve cycle efficiencyis the use of reheat. This variation involves expanding steam in thesteam turbine from design inlet conditions down to some specified reheatpressure. At this point, some energy has already been extracted from thesteam to produce shaft horsepower. This lower energy content steam isthen redirected to the boiler where it is reheated to a highertemperature. This higher energy content steam is then sent back to thesteam turbine to produce more power. More than one reheat can beutilized in the cycle. Again, for the given design conditions, inletpressures, inlet temperatures, and other conditions, the reheat isdesigned for the greatest benefit and increase in cycle efficiency.

Other Factors

Other factors that affect cycle efficiency include inlet steam pressure,inlet steam temperature, and exhaust pressure. Typically, higher inletpressures and higher inlet temperatures yield higher cycle efficiencies.Lower exhaust pressures typically also yield higher cycle efficiencies.Exhaust pressures are normally limited by ambient factors, such as thetemperature of the river water, ambient air, or other fluid used to coolthe condenser. This will set the limit for the exhaust pressure, and thecondenser and associated equipment will be designed to approach thislimit, based upon evaluated parameters such as size, cooling mediumavailable, environmental factors, and cost.

Design Limitations

Inlet pressure and inlet temperature are typically selected by the plantdesign engineer. However, there are limits that are imposed in thesedesigns. As the inlet pressures are increased, the stresses on theboiler tubes, steam turbine casing, and steam turbine internals areincreased. These stresses impose limits on the manufacturer's ability toproduce this equipment, or economic limitations on the feasibility ofproducing this equipment. In addition, above 3206 psia, steam no longercan coexist as both water and steam. This point is referred to as thecritical point of steam, and above this pressure steam does not boil.Instead, both water and steam are a fluid and a more intricatesuper-critical boiler is required to produce steam above this pressure.At higher temperatures, the allowable stress of the boiler tubes, steamturbine casing, and steam turbine internals is reduced, and near thecurrent limits, conventional steam turbine materials rapidly loose theirproperties as the temperature is increased only small amounts (50° F.).Conventional large steam turbines built as state of the art machineshave HP inlet temperature limits in the range of 1050° F.

Steam Cycle Optimization

Once a boiler steam pressure and temperature is selected, the steamcycle then must be optimized. A typical high efficiency steam cycle willinvolve the use of feedwater heaters, a reheater, a reheat steamturbine, boiler feed pumps, and a condenser. A descriptive document oncycle optimization is an informative paper issued by General ElectricCompany (GE) entitled “Steam Turbine Cycle Optimization, Evaluation, andPerformance Testing Considerations” (General Electric ReferenceGER-3642E, 1996) by James S. Wright. This document provides relativeperformance variations for different cycle parameters such as pressure,temperature, number of reheats, and number of feedwater heaters.

Rankine Cycle Example

FIG. 3 is a schematic representation of a Rankine Cycle with bothfeedwater heating and reheat. This sub-critical Rankine Cycle works byproviding water to the inlet of the boiler feed pump (BFP) (301). Thewater is then pumped to a desired discharge pressure by the BFP (301).This pressurized water is then sent through the feedwater line (302) tofeedwater heater (FWH) (303) and through line (304) to feedwater heater(305). The feedwater heaters (303, 305) preheat the feedwater before itenters the boiler at the boiler inlet (306). This preheated feedwatertravels to the evaporator section (307) of the boiler where heat isadded to the pressurized water.

Steam exits the boiler section at (308) and continues to superheatersection (309) and exits at (310). This superheated steam is sent to thehigh-pressure (HP) section of the steam turbine (311). The steam expandsthrough the HP section to (312), and then returns to the reheat sectionof the boiler (RHT) (313) where heat is added to return the steamtypically to a temperature at or near the inlet steam temperature. Thisreheat steam is then sent to the Intermediate Pressure (IP) section ofthe steam turbine at (314). This steam then expands through the IPturbine section (315) and produces shaft horsepower. The steam thenexits the IP section and via the crossover pipe (316) and goes to the LPsection of the steam turbine (317).

Due to the high volume flows at low-pressure, the LP section istypically a double flow section on large units, so steam enters themiddle of the casing and travels both forward and aft through theblading to produce more shaft horsepower. The steam then exhausts at(318) into the condenser (COND) (319). Condensed steam leaves thehotwell (330) and returns via the feedwater line (320) to the inlet ofthe BFP (301). For feedwater heating, steam is extracted from the IP andLP sections of the steam turbine at (321) and (324) and sent tofeedwater heaters (305) and (303) respectively via lines (323) and(326). Non-return valves are used in these lines, (322) and (325), toprevent backflow of steam to the ST in case of a trip (emergencyshutdown) condition when pressures in the turbine will rapidly drop tocondenser pressure. These valves are safety devices only, and are eitheropen or closed. Steam from these extraction lines preheats the feedwateron its way to the boiler. The steam from the extraction lines iscondensed in the feedwater heaters and the condensate (327, 328) isreturned to the inlet of the BFP (301). Again, shaft horsepower producedin the ST is converted into electrical power in the generator (GEN)(329).

For larger, central power plant applications, typical inlet pressuresfor sub-critical applications are 1800 and 2400 pounds per square inchgauge (psig). For supercritical applications, pressures of 3500 psig andgreater are employed. Inlet steam temperatures for most large steamturbines are limited to about 1050° F. for both the inlet and reheatsteam. However, some advanced technology steam turbines are utilizinginlet temperatures of 1070° F. for the HP inlet and 1112° F. for reheat,as detailed in a descriptive document on steam turbines issued byGeneral Electric Company (GE) entitled “Steam Turbines forUltrasupercritical Power Plants” by Klaus M. Retzlaff and W. AnthonyRuegger (General Electric Reference GER-3945, 1996).

Rankine Cycle Efficiency Comparison

Based upon a steam turbine with a 90% efficiency, FIG. 4 illustrates arelative comparison of a basic Rankine Cycle (Option 1), (excludingboiler efficiency and parasitic power requirements) to one that usesonly reheat (Option 2, Option 3), and to one that uses both reheat andfeedwater heating (Option 4, Option 5). Variations in the inlet pressurewith reheat (Option 3) and feedwater heating (Option 5) are alsoincluded. Option 6 and Option 7 are for supercritical steamapplications. Option 6 is a supercritical steam cycle withultrasupercritical (inlet or reheat temperatures above 1050° F.) steamconditions and double reheat (steam is reheated twice, at two separatepressure levels, in the boiler). Option 7 is the same as Option 6 withthe addition of feedwater heating. For the purposes of this comparison,only two extractions were utilized and the extraction pressures wereassumed to be at the cold reheat pressure and the crossover pressure(2nd cold reheat for supercritical applications). More feedwater heaterswill yield even better cycle efficiencies. General Electric Company (GE)informative document entitled “Steam Turbine Cycle Optimization,Evaluation, and Performance Testing Considerations” (General ElectricReference GER-3642E, 1996) by James S. Wright provides data for theselection of the optimum number of feedwater heaters, stating that a1.5% heat rate penalty is assessed for only three feedwater heatersversus seven. Therefore, the feedwater heating cycle efficiency shown onFIG. 4 (Options 4, 5, and 7) has room for improvement. With reheat,optimum feedwater heating, and ultrasupercritical steam conditions,overall plant cycle efficiencies in excess of 45% are possible.

The overall plant cycle efficiency includes not only the basic steamcycle efficiency as shown in FIG. 4, but also the boiler efficiency andparasitic power requirements such as the boiler feed pumps and thecondenser circulating water pumps. As stated in POWER MAGAZINE, (ISSN0032-5929, July/August 1998, page 26):

“Over the last few years, new designs have evolved to boost efficienciesof steam power plants, and the steam turbine is a large part of thiseffort. Efficiencies of 45% (LHV) [Lower Heating Value] or higher arenow possible with the latest fossil-fired steam plants using the higheststeam parameters, advanced feedwater heating cycles, boiler and turbinemetallurgies, etc.”

To obtain an overall plant efficiency of 45% LHV, including the boilerefficiency and parasitic power requirements, typically means that thebasic steam cycle efficiency must be even higher than 45%. With a boilerefficiency of 85%, parasitic power requirements of 2.5%, a ratio of HHV(higher heating value) to LHV (lower heating value) of fuel of 1.11(typical for natural gas), and a plant efficiency of 45% (LHV), thebasic steam cycle efficiency would calculate to

48.9%=0.45/(0.85×(1−0.025)×1.11)  (1)

As seen from FIG. 4, the use of a reheat steam cycle can increase thebasic Rankine Cycle efficiency by 4.79% at the tabulated steampressures. However, the use of reheat as well as increased inletpressures and feedwater heating can boost efficiency by at least 10.3%for sub-critical steam conditions. (Note that efficiency improvement isthe ratio of a particular option efficiency to the base efficiency.Thus, a 40% efficient cycle would convert 40% of the input energy toelectricity. That is twice as much as a 20% efficient cycle. Therefore,the efficiency improvement from a 20% efficient cycle to a 40% efficientcycle is 100%, or twice as much output).

Fuel efficiency is of the utmost importance at power plants and a largecentral coal-fired power plant may expend approximately US$140 millionannually for fuel, assuming a plant rating of 1000 MW, 45% thermalefficiency LHV (lower heating value of the fuel), US$2.00 per millionBTU for fuel, and 8500 operating hours per year. Given these facts, evena 1% increase in efficiency will equate to large cost savings in fuel(US$1.4 million annually).

Combined Cycle Application

Although the Rankine Cycle has been well proven, today's more strictenergy and environmental standards require more emphasis be placed onfuel efficiency and low emissions from power plants. As a result, newcombined cycle plants are being designed and built.

FIG. 5 is a conceptual schematic for a combined cycle application. Inthe general sense, combined cycle is not limited to a Brayton Cycletopping cycle and a Rankine Cycle bottoming cycle, but can be anycombination of cycles. The topping and bottoming cycles could be thesame cycle using different fluids. Either way, FIG. 5 would beapplicable. In FIG. 5, the topping cycle fluid (TCF) (501) enters thetopping cycle engine (TCE) (502) where fuel (CFT) (503) is added toraise its temperature. The fluid performs work that is converted by thetopping cycle engine into shaft horsepower. This shaft horsepower drivesthe topping cycle load (TCL) (504). This load could be an electricalgenerator, pump, compressor, or other device that requires shafthorsepower. The exhausted fluid from the topping cycle engine isdirected through an exhaust line (505) to a heat recovery device (HRD)(506), and then exhausts to an open reservoir (507).

For this example, the topping cycle is an open cycle. In other words,the topping cycle fluid is taken from a large reservoir and dischargesto that same reservoir. The heat recovery device (506) captures aportion of the topping cycle exhaust energy and transfers it to thebottoming cycle fluid (BCF) (508). In this example, the bottoming cyclefluid is heated at three separate pressure levels: a high-pressure line(509), intermediate pressure line (510), and low-pressure line (511).These fluids then travel to the bottoming cycle engine (BCE) (512) whereit produces shaft horsepower to drive the bottoming cycle load (BCL)(513). Again, this load could be an electrical generator, pump,compressor, or other device that requires shaft horsepower.

From the bottoming cycle engine, the bottoming cycle fluid enters a heatexchanger (HEX) (514) where heat is rejected. The bottoming cycle fluidthen enters a pump or compressor or other fluid transfer device (FTD)(515) where it is then returned to the heat recovery device (506). Forthis example, the bottoming cycle is a closed cycle, meaning that thebottoming cycle fluid is continuously circulated within a closed loop.There could be more than two cycles in this process, and any of thecycles could be either open or closed loop. This describes the basicfundamentals of a combined cycle application.

HRSG in Combined Cycles

In many cogeneration and combination GT/ST power plants built today,combined cycle plants have come to mean power plants that utilize aBrayton Cycle as the topping cycle and a Rankine Cycle as the bottomingcycle. These plants utilize a gas or combustion turbine (GT) as theprime mover (Brayton Cycle machine), with a boiler at the exhaust of thegas turbine to recover the waste heat. This boiler is typically referredto as either a waste heat boiler (WHB) or a heat recovery steamgenerator (HRSG). It may also have burners in place to increase theexhaust gas temperature and produce more steam than that available fromjust the waste heat (supplemental firing). The HRSG produces steam thatis then sent to the steam turbine (ST) to produce more power. Due to thehigh temperatures of the working fluid in the GT (approximately 2400° F.for GE industry standard “F”-class technology machines and 2600° F. forWestinghouse industry standard “G”-class technology machines), andrecovery of the waste heat, the combined cycle plants are much more fuelefficient than the conventional steam plants. In addition, with advancesin GT technology and the use of either distillate oil or natural gasfuel, the emissions from the combined cycle plants are extremely low.FIG. 6 illustrates a typical combined cycle application.

The HRSG is distinctly different from a conventional Rankine Cycleboiler. A Rankine Cycle boiler is fueled by a variety of fuels,including oil, natural gas, coal, biomass, as well as others. TheseRankine Cycle boilers may also use a combination of fuels as well. TheHRSG may not utilize any fuels at all, but only capture and utilize theexhaust heat from the GT. If it is supplementary fired, the HRSG willrequire more refined fuels such as natural gas or distillate oil. Solidfuels such as coal and biomass are not typically utilized in these typesof boilers.

As seen from FIG. 6, there are numerous sections to the HRSG, includingthree evaporator sections (one for each pressure level), economizers,superheaters, and a reheater. Sections (601) and (602) are economizers.These are large tubed sections in the HRSG that preheat water before itis converted into steam in the Evaporator. Sections (603), (606), and(609) are LP, IP, and HP evaporators respectively. Sections (604),(605), and (607) are feedwater heaters. Section (608) is the IPsuperheater while sections (610) and (612) are HP superheaters. Section(611) is the reheater section. These HRSGs are typically very large andheavy pieces of equipment with literally miles of tubes inside.

Steam from each pressure level is utilized in the power plant whererequired, but essentially, most steam is generated for the purpose ofproducing additional power in the ST. This means that the lower pressurelevels of steam must be introduced or admitted to the ST at the properpoint on the ST other than the HP inlet. It also means that the ST musthave provisions (openings, nozzles, connections, trip valves, etc.)where this steam may be admitted, and that at the operating conditionsthe steam pressure in the ST at these connections must be less than thepressure of the steam from the HRSG corresponding boiler sections.Otherwise, steam will not flow into the ST.

As noticed from a comparison of FIG. 6 with FIG. 3, the conventionalRankine Cycle utilizes feedwater heaters that take steam from the ST topreheat feedwater, while the HRSG utilizes the GT exhaust heat toprovide this function. Therefore, conventional steam fed feedwaterheaters are not typically employed in combined cycle applications. In GEinformative document GER-3582E (1996), entitled “Steam Turbines forSTAG™ Combined Cycle Power Systems”, M. Boss confirms that feedwaterheaters are not utilized in the prior art:

“Exhaust sizing considerations are critical for any steam turbine, butparticularly so for combined-cycle applications. There are usually noextractions from the steam turbine, since feedwater heating is generallyaccomplished within the HRSG”.

Another modification typically used for combined cycle applications isthe use of two boiler feed pumps (630), and (631), typically referred toas the LP and HP BFPs respectively. This arrangement allows the LP pumpto provide pressurized water for the LP and IP pressure levels and theHP pump provides water for the HP pressure level, which saves pumphorsepower. For large combined cycle applications, the steamturbine/condenser arrangement is similar to the Rankine Cycle depictedin FIG. 3, (although internally, the steam path designs are totallydissimilar).

HRSG/Combined Cycle Disadvantages

General Disadvantages

With current technology, maximum inlet pressures to the steam turbinefor combined cycle applications are nominally 1800 psia with inlet steamtemperatures near the limit of 1050° F. for both the inlet and reheatsteam. Some of the disadvantages of this HRSG arrangement for combinedcycle applications are as follows:

1. Steam cycle efficiencies are much lower than those of conventionalsteam power plants.

2. Multiple evaporator sections are required to maximize heat recovery.This results in increased equipment and maintenance costs.

3. Multiple evaporator sections require the plant operators and controlsystems to monitor and control all boiler (evaporator) drum levels.

4. The HRSGs with the multiple sections are very large, require largeamounts of infrastructure building volume, large amounts of floor space,and large foundations to support the weight of the HRSG.

5. The HRSGs are expensive (approximately $10 million for a HRSG thatrecovers exhaust gas heat from one GE Frame 7 GT).

6. Maintenance increases with the number of components, evaporatorsections, controls, and other devices.

7. Low-pressure steam (steam other than the highest pressure steam) hasmuch less ability to produce power in the ST than higher pressure steam.

8. Partial load, off design operation, and other conditions besides thedesign conditions typically have reduced heat recovery and lower cycleefficiencies.

9. Increased amounts of tubing in the HRSG to enhance heat recovery addflow restriction to the exhaust gases from the GT and this increasedback pressure decreases GT output and efficiency.

10. Gas turbine exhaust temperatures are not sufficient to produce someof the elevated steam conditions now used in advanced steam cycles (600°C. which is equivalent to 1112° F.).

11. Balancing problems in the reheat lines with multiple GTs (typicallythree or more) make it difficult to utilize large STs in combined cyclepower plants in the prior art. For modern, large, and efficient combinedcycle plants such as a GE S207FA, the steam turbine rating isapproximately 190 MW, which is much smaller than GE's large steamturbines which can exceed 1200 MW. For more information on large steamturbines, reference the informative paper issued by General ElectricCompany (GE) entitled “Steam Turbines for Large Power Applications” byJohn K. Reinker and Paul B. Mason (General Electric Reference GER-3646D,1996).

Part Load Operation Inefficiencies

Another disadvantage of the combined cycle application is partial load(part load) operation. As the system to which a power plant is connectedreduces its load requirement, the power plant must respond by providingless output. This load modulation allows for a constant speed on themachinery and a constant frequency of power (e.g., 60 Hz in the UnitedStates and 50 Hz in Europe). To modulate the load at a combined cycleplant, less fuel is burned in the GT, and the power output is reduced.This typically requires a reduction in the GT firing temperature and/ora reduction in GT airflow.

Part load operation reduces the efficiency of the GT, thus reducing theefficiency of the entire combined cycle plant. FIG. 7 illustrates atypical curve for a large modern GT with inlet guide vanes (IGVs) tomodulate inlet airflow. Even with the enhanced part load efficiencygained by the use of IGVs, at 60% load (Generator Output—PercentDesign), the GT consumes over 70% of the fuel required at full load(Heat Consumption—Percent Design). This represents a 17.5% increase inheat rate (specific fuel consumption). For GTs without IGVs, this decayin performance would be even more pronounced.

To help offset this part load decay, plus provide more power output fora given amount of hardware (sometimes referred to as power density),manufacturers can provide combined cycle power plants with two GTs, eachwith its own HRSG, feeding into one ST (referred to as a 2-on-1arrangement). With an arrangement such as this, when the power plantload decreases to slightly less than 50% for a 2-on-1 arrangement(2-GTs, 1-ST), one GT can be shut down, and the remaining GT can returnto near 100% output. This mode of operation increases part loadefficiency below 50% of total plant load as illustrated graphically inFIG. 8. This graphically illustrates a typical two GT comparison takenfrom GE informative document GER-3574F (1996), entitled “GECombined-Cycle Product Line and Performance” by David L. Chase, Leroy O.Tomlinson, Thomas L. Davidson, Raub W. Smith, and Chris E. Maslak for acurve of GE combined cycle part load performance with a 2-on-1arrangement. For a 3-on-1 arrangement, switchover from three to two GTscould occur slightly below 67% load. This still provides for substantialincrease in plant heat rate at part load conditions. Note that providingthis increase in part load efficiency occurs as a result of higherequipment costs. The prior art has yet to solve the efficiency problemwithout the addition of more equipment that increases the overall powerplant costs.

Supplementary Firing of HRSG

Another solution to add flexibility to the operation of a combined cyclepower plant is the use of supplementary firing in the HRSG. This mode ofoperation is when fuel is burned in the HRSG just after the GT (or atsome intermediate point within the HRSG). This increases the temperatureof the exhaust gas to the HRSG and produces more steam that can be sentto the ST. This allows the plant to produce more power. However, theplant heat rate increases, and fuel efficiency decreases accordingly.This result is stated by Moore of GE in U.S. Pat. No. 5,649,416. Thispatent, as well as U.S. Pat. No. 5,428,950 by Tomlinson, is referencedby Rice in U.S. Pat. No. 5,628,183. Therefore, supplementary firing ofthe HRSG is considered by the manufacturers to be a means to obtain moreoutput, but with a penalty on efficiency. GE informative documentGER-3574F (1996) entitled “GE Combined-Cycle Product Line andPerformance” by David L. Chase, Leroy O. Tomlinson, Thomas L. Davidson,Raub W. Smith, and Chris E. Maslak states that

“incremental efficiency for power produced by supplemental firing is inthe 34-36% range based upon lower heating value (LHV) of the fuel.”

Also in this GE document, Table 14 indicates that HRSG supplementalfiring can increase combined cycle plant output in the prior art by 28%,but only with an increase in overall combined cycle heat rate (specificfuel consumption) of 9%. No technique has been shown in the prior art toeliminate this heat rate penalty associated with supplemental firing.

Additionally, supplemental firing in the prior art can be utilized toachieve higher ST/GT ratios than is typical for conventional combinedcycles. However, operation at these high levels of ST/GT output aretypically short in duration to meet peak power demands, and long termoperation at these ratios is not economical. Therefore, conventionalcombined cycle power plants that are designed with ST/GT ratiosapproaching unity do not operate predominantly as Rankine cycle powerplants, but do so only to satisfy temporary peak plant loads, and do sowith a significant efficiency penalty at all operating conditions.

Gas Turbine Performance Decay

As mentioned in the discussion on the Brayton cycle, approximately ⅔ ofthe energy produced by the turbine section of the gas turbine isrequired to drive the compressor section, with the remaining ⅓ availableto drive a load. This power consumed by the compressor at 67% of theturbine output, is much higher than the Rankine cycle example where theboiler feed pumps (BFP) only consumed 2% of the turbine power.Therefore, the GT is susceptible to performance decay if the compressordoes not maintain optimum efficiency.

For example, a typical efficiency for an axial flow air compressor usedwith a large GT might be 90%. Therefore, if the compressor requires 67%of the turbine section output, the ideal power (100% efficient) wouldonly be (0.67*0.90)=0.603 or 60.3%. If the compressor efficiency were todecay by 2.5%, its new efficiency would be (0.90*0.975)=0.8775 or87.75%. The compressor power required would now be (0.603/0.8775)=0.6872or 68.72%. Turbine net output would be reduced from 33% (1.00-0.67) to0.3128% (1.00-0.6872). This represents a 5.2% loss in output(0.3128/0.33=0.9479). Therefore, it can be readily seen that smalldecreases in efficiency for the GT compressor lead to large decreases inefficiency and output for a GT.

The efficiency and rating loss of 5% from the above example is typicalof many GTs after about one or two years of operation. This efficiencydecay is largely a result of worn clearances in the compressor anderosion of the compressor blade tips. New blades and seals willtypically restore the compressor efficiency to almost “new” conditionefficiency. However, this is a costly and time consuming repair, andwould probably only be done at major inspections, which are scheduledapproximately every four years for modern GTs. Therefore, plant ownersand operators will need to plan on this performance decay between majoroverhauls of the GTs.

Candidates for Improvement in the Prior Art

From the foregoing discussion it can be seen that parameters of thecurrent and defined technology that are candidates for improvement maybe described as follows:

Flexibility

Due to the electrical load demand in a particular region or marketplace,the electric utility (which distributes electrical power to the endusers) determines the need for power based on current demand and futureprojections. For example, if this load was determined to be 850 MW, in aconventional Rankine cycle configuration the utility/Power Developerwould contract with an Architect/Engineering (AE) firm to design andbuild such a plant. The boiler, pumps, condenser, steam turbine, and allthe other plant auxiliaries would then be designed for the specifiedoutput of 850 MW. This can be accomplished largely due to the fact thatsteam turbines are custom designed and manufactured. However, with gasturbines being production line items, and combined cycles beingprimarily gas turbine based power plants, to achieve the highestefficiencies and best capital cost, a utility and/or power developer canno longer specify just their plant output, but must find the best fitfor their needs from the available combined cycle offerings from thevarious manufacturers. For example, a review of the available combinedcycle plants from the 1997 TURBOMACHINERY HANDBOOK, (USPS 871-500, ISSN0149-4147), indicates that there are no 850 MW combined cycle plantsavailable for 60 HZ applications. Thus, a plant developer's designflexibility is constrained by the current state of the art of combinedcycle power plant equipment. This implies that in certain circumstancesthe equipment complement for a given power plant installation will notbe optimal because of constraints placed on plant equipmentconfigurations by the current state of the art.

Efficiency

Combined cycle power plants are extremely energy efficient compared toother conventional means of producing electricity. However, a largecentral combined cycle power plant rated for 1000 MW at 55% thermalefficiency LHV (lower heating value of the fuel) operating 8500 hoursper year at full load with a fuel cost of US$3.00 per million BTU offuel will expend approximately US$175 million annually for fuel. Even a1% increase in efficiency will equate to large savings in fuel (US$1.75million annually).

In U.S. Pat. No. 4,333,310 issued to Robert Uram, a control method isutilized which monitors the steam temperature to the ST and modulatesthe afterburner (supplemental firing) to control the temperature of thesuperheated steam. While providing optimum ST inlet temperatures, thisfunction does little to affect load. In this patent, Uram states

“It is desired that the steam turbine be operated in what is called a“turbine following” mode wherein the plant is supplying electrical powerto a load, such that the steam turbine follows the gas turbines and eachafterburner positively follows a respective gas turbine. In other words,the heat contributed by the afterburner follows the temperature of thegas turbine exhaust gas, and the steam produced by the gases exhaustedfrom the afterburners is used in total by the steam turbine.”

These teachings of the prior art are in direct contrast to that of thepresent invention in which the heat contribution via supplemental firingis independent of the gas turbines, and the gas turbines are designed tooperate substantially at their optimal full rated capacity.

Installed Cost

Next to fuel costs, the largest cost for a combined cycle plant istypically debt service. Manufacturers, engineering firms, and owners arealways interested in finding ways to reduce the installed cost of powerplants. At 8% interest and US$450 per kW of capacity, a 1000 MW combinedcycle power plant would have a debt service of approximately US$45million per annum for 20 years. Reducing the capacity cost, in US$/kW,directly reduces the debt service.

Temporary Capacity Extension During Peak Demand Loading

One dilemma that faces power plant owners and utilities is the properselection of power plant capacity. Selecting a plant that is too smallresults in power shortages, brownouts, and/or the need to purchaseexpensive power from other producers. Selecting a plant that is toolarge results in operation at lower efficiency during part load andincreased capital cost per kWh produced. In many situations the problemfaced by power plant developers is the need to provide for peak powerneeds and temporary demand loading. This peak may occur only in certainseasons for a limited span of time. Typically in the summer monthsduring peak hours on the hottest days is the most challenging time forpower producers to meet the system load. Having the ability to provideexcess capacity during this time period is highly desirable, and in theemerging arena of electrical power deregulation, it may prove to be verylucrative. For example, in the early summer of 1999, power shortages inthe Northeast United States have caused concern for the system's abilityto meet peak power demands. Some local newscasts have reported costs forcapacity at $30/MWh during normal periods and as high as $500/MWh duringpeak. However, even much greater capacity costs have been incurred, asreported in POWER MAGAZINE, (ISSN 0032-5929, March/April 1999, page 14):“Reserve margins are down nationwide from 27% in 1992 to 12% in 1998,according to Edison Electric Institute, Washington, D.C., becausederegulation uncertainty has caused capacity additions to stall. Lastsummer's Midwest [United States] price spikes, up to $7000/MWh, garneredmost of the press coverage, but spikes of $6000/MWh also occurred inAlberta . . . ”

However, providing peak power will not be lucrative if the power plantowners have to pay for this capacity, pay the debt service, and yet makerevenue on this extra capacity only during a few days of the year.Therefore, power plants that can provide more output than normal duringpeak demand hours are needed to help supply system load during thesepeak demands.

Reference FIG. 31B for a graphic illustrating the relative percentage oftime that a typical power plant spends in peak, intermediate, and baseloading conditions. From this graphic it can be surmised that it wouldnever be profitable to design a power plant to peak loading conditions,as they occur less than 10% of the time. Since prior art power plantsare generally incapable of wide variations in peak power output, theonly practical option available for present power providers is topurchase power over the electrical grid during times of peak powerdemand. The present invention teaches a system and method which permitsthis peak demand to be satisfied without the need for purchasingexternal power over the electrical grid, thus providing an economicadvantage over the prior art.

Non-Local Power Generation/Distribution Reliability Issues

One significant problem with the prior art is that the plant capacity isin general a relatively fixed and narrow range of power generationoperation. When peak power demands are placed on the electrical grid,electrical power must be purchased from elsewhere on the grid whereelectrical demand relative to remote plant capacity is lower. There areseveral major problems with this mode of providing for peak power byrerouting remotely generated power plant capacity.

First, there exist losses associated with transmission of power fromremote sites to the place where the electrical power is being demanded.For example, a hot summer day in New York City may require diversion ofpower from Canada or the western United States, resulting in significantline losses during transmission.

Second, there is a reliability drawback in purchasing power from distantparts of the grid during periods of peak load. While it is possible toredistribute power, the tradeoff is instability in the electrical grid.What can happen is that small failures in remote parts of the grid cancascade throughout the grid to either cause additional equipmentfailures or cause instability in the grid voltage. Thus, whilepurchasing power from remote power plants may alleviate some localreliability problems with respect to providing electric power, thetradeoff is an overall reduction in the reliability of the entireelectrical grid. Thus, relatively insignificant events in remote partsof the country can cascade throughout the electrical grid and result inserious electrical failures in major metropolitan areas.

Thus, given the above reliability concerns, it is in general alwaysbetter to be able to provide electrical power local to the demand forthat power. While the existing prior art relies heavily on power sharingand distribution, the present invention opts for the more reliablemethod of generating the power locally to provide a power generationsystem that is more efficient and reliable that the current state of theart. It is significant to note that the prior art limitations on plantoutput during peak load generally preclude local generation of therequired peak power demand. This forces traditional power plants topurchase power from remote power plants at a substantial (10× to 250×)price penalty.

Operation and Maintenance Costs

Costs for personnel, fuel, maintenance, water, chemicals, spare parts,and other consumables, including other costs such as taxes andinsurance, all contribute to Operation and Maintenance (O&M) costs. Asthe plant size grows, the amount of equipment increases, and as thecomplexity of the equipment increases, O&M costs also increase. In thequest for higher efficiency, more elaborate and expensive technology isbeing utilized in the gas turbines. The maintenance costs associatedwith exotic new materials, intricate blades, and complex hardware isprojected to be significantly more expensive than the slightly lessefficient, proven gas turbine hardware and associated plant designs.

To be prepared for an equipment failure, plant owners must retain largequantities of spares on hand at their facility. This constitutesinventory that has high costs in terms of both unused capital and taxes.Methods to reduce O&M costs are always desired by the plant owners andoperators.

Fuel Gas Compression

Current projections are that natural gas will have a stable supply andprice structure until the year 2010. This fuel is clean, efficient, andinexpensive, and thus is the preferred fuel for combined cycleapplications. However, if the power plants are not located in closeproximity to major natural gas pipelines, the lower pressure natural gasmay have to be compressed to a sufficient pressure to be used in the GT.In addition, the higher efficiency GTs such as the Westinghouse model501G require higher fuel gas pressure than GTs with lower pressureratios, such as a GE model PG7241FA GT. This need for higher pressurenatural gas requires expensive natural gas compressors that are criticalservice items (the plant cannot operate without them). These natural gascompressors require frequent maintenance and also consume parasiticpower (the power to run the compressors reduces the net power availablefrom the power plant to the grid). Reducing the need for thesecomponents reduces the plant installed cost, reduces real estaterequirements, improves reliability, and increases the plant net output.

Plant Reliability

Electrical power reliability has become a facet that is demanded by boththe residential consumer and industrial user of electricity. Therefore,the technology to produce power must be proven and reliable. In U.S.Pat. No. 5,628,183, Rice proposes a higher efficiency combined cyclepower plant. However, this system requires the use of diverters in theHRSG, natural gas reformers, and the use of steam superheated to 1400°F. These systems will all add greatly to the installed cost and O&Mcosts. In addition, to date, boiler tubes, HRSGs and STs have notdemonstrated long term reliable operation at elevated temperatures above1150° F., and HRSGs with diverters and natural gas reformers are as yetunproven in the marketplace.

Air Consumption

GT engines consume large quantities of air. A typical combined cycleinstallation will consume approximately 20 lbs. of air per kW ofelectricity produced. This equates to approximately 260 cubic feet (atsea level) per kW. This air must be filtered before it enters the GT toprevent foreign object damage in the GT. Periodically, the air filtersmust be cleaned and/or replaced. This adds to the O&M costs andincreases plant downtime (time when the plant is out of service andunavailable to produce power).

In addition, the air consumed by the GT is discharged to the HRSG andthen exhausted to atmosphere. As more air is consumed, more air must beexhausted. This represents an efficiency loss as the HRSG exhausttemperature is typically about 180° F. In addition, this airflow servesto heat the atmosphere and contribute to local air quality problems.

Plant Emissions

In order to obtain a permit to operate, a power plant must first obtainan air permit. This permit typically states the allowable levels ofcertain criteria pollutants that a plant may emit. Combined cycle powerplants are very clean producers of power compared to other conventionalmethods, but are typically plagued by one criteria pollutant, nitrousoxides (NOX). This criteria pollutant is usually controlled by steamand/or water injection into the GT, dry low NOX combustion systems,and/or exhaust gas aftertreatment. The exhaust gas aftertreatmenttypically employed is “Selective Catalytic Reduction” (SCR) whichessentially works by injecting ammonia (NH.sub.3) into the exhaust gasstream in the presence of a catalyst at a specified temperature range toreturn the NOX formed by the combustion process into N2 and H₂O.

In U.S. Pat. No. 3,879,616 by Baker, et al., U.S. Pat. No. 4,578,944 byMartens et al., and U.S. Pat. No. 5,269,130 by Finckh, et al., the plantload is controlled by changes in the GT output. However, at partialload, GT NOX emissions are typically increased. Therefore, it may benecessary to introduce more ammonia into the exhaust gases for emissionreduction. This increases O&M costs, and can be significant to the pointwhere, at the plant design stage, the desired GTs cannot be used due tohigh emission levels at part load operation. Also, if run at full load,some plants may not require SCR, but due to part load operation, SCRwill be required. Another factor related to emissions is airconsumption. GTs require large amounts of air, and the more air that isconsumed, the more potential there is for emissions.

Environmental Considerations

Besides air emissions, a power plant must be concerned with otherenvironmental impacts as well. To operate a steam plant, a clean sourceof water must be available to provide make-up water. This make-up wateris used to replace steam/water that is lost to ambient through leaks,blowdown, or other loss. Blowdown is the water that is taken from theevaporator sections of the HRSG and dumped to the sewer. This blowdowntypically is taken from a low point on the HRSG to remove feedwater thathas high concentrations of minerals and deposits. This process helpskeep the steam path clean and minimizes ST deposits and blade failuredue to stress corrosion cracking. This blowdown must be discharged intorivers, streams, etc. and as such requires water permits that may bedifficult and time consuming to obtain from regulatory authorities.

Distributed Plant Control System (DCS)

Modern combined cycle plants typically use a distributed control system(DCS) to control the entire plant. These DCS controls integrate with theindividual control systems on the GTs and STs. Many other parameters canbe monitored and controlled by the DCS. Use of controls to better eitherefficiency or operation is described in U.S. Pat. No. 3,879,616 by Bakeret al., U.S. Pat. No. 4,201,924 by Uram, and U.S. Pat. No. 4,578,944 byMartens et al. None of these patents, however, provide control of heattransfer in the HRSG. In U.S. Pat. No. 5,269,130 by Finckh et al., amethod of controlling excess heat in the HRSG is utilized for part loadoperation of the GT. This method, however, does not providecomprehensive control, but only a means for recovering low temperaturewaste heat. None of the aforementioned patents has devised a method tocontrol the exhaust gas temperature of the HRSG to its optimumtemperature.

Plant Operational Efficiency

Combined cycle power plants in the prior art that are designed formaximum efficiency typically utilize multi-pressure HRSGs, commonly atthree pressure levels. For each HRSG, and for each pressure level, theoperations staff must monitor the steam drum level. Also, parameterssuch as water quality and chemical content must be monitored for eachHRSG. Since the system load for any utility is constantly changing,combined cycle power plants are required, like other power producingplants, to be dispatched, or provide load as required to the electricalgrid. This means the power plant will not operate at a fixed load, butwill constantly be modulating load to meet the system demand. Toincrease load, supplementary firing (additional fuel burned at or nearthe inlet to the HRSG to add energy to the exhaust gases) can beaccomplished. However, this is detrimental to overall plant efficiency.This is noted by Rice in patent U.S. Pat. No. 5,628,183 with referencesto Westinghouse and General Electric studies. Moore in patent U.S. Pat.No. 5,649,416 states that

“Supplemental firing of the heat recovery steam generator can increasetotal power output and the portion of the total power produced by thesteam turbine, but only with a reduction in overall plant thermalefficiency.”

Therefore, it is common in combined cycle plants to see little or nosupplemental firing used. Therefore, to change and meet varying systemloads, the GTs are brought from full load to part load operation.

As well as increasing emission levels as previously mentioned, this partload operation also has a detrimental effect on efficiency. FIG. 7 is arepresentative curve of GT efficiency versus load. At 100% load itconsumes 100% fuel, however, at 60% load, it consumes 70.5% of full loadfuel. This is an increase of 17.5% in specific fuel consumption. Forlarge central power plants, this factor equates to significant addedfuel costs. In addition, operation at part load on the GT typicallyincreases the emission levels for the most difficult criteria pollutant,NOX. Part load operation of the GT also changes the exhaust gas flowthrough the HRSG. This change in flow upsets the heat transfer in theHRSG since this device is constructed with fixed heat exchange surfacearea. This phenomenon, as well as reduced GT efficiency, contributes topoorer overall efficiency at part load operation. If part load operationchanges temperatures in the HRSG significantly, this could lead toineffective operation of the SCR.

Steam Turbine Exhaust End Loading

Besides inlet pressure and temperature limitations, another commonlimitation for the steam turbine (ST) is the exhaust end loading. Thisessentially is a function of two parameters, exhaust end flow andexhaust pressure. These two factors essentially determine the volumetricflow through the last stage blading of the ST. For optimum operation,there is a range of volumetric flow typically specified by the STmanufacturers. As this volumetric flow increases, larger blades and/ormore exhaust sections may be required.

However, due to mechanical limitations (centrifugal force), once thelargest available blade volumetric limits are reached, more sections andmore blades must be added to the exhaust end of the ST to accommodatethis flow. This adds to the installed cost and increases the real estaterequirements of the ST. Due to its configuration, a conventionalcombined cycle sends HP steam to the ST HP inlet, then adds steam fromthe IP section of the HRSG to this flow at the ST IP section inlet, thenadds more steam from the LP section of the HRSG to this flow at the STLP section inlet.

Therefore, in this arrangement, the HP and IP sections of the ST seerelatively lower flows and lower volumetric efficiencies than the LPsection. This arrangement leads to STs that are at or near the exhaustend limit of the ST. This provides for little in the way of temporarycapacity extension for peak power production and leaves little or noability to uprate (increase) the ST in the future to a higher powerrating. Overall, this ST arrangement is less efficient than conventionalsteam plant STs since the HP and IP sections have low volumetric flows.

In GE informative document GER-3582E (1996), entitled “Steam Turbinesfor STAG™ Combined Cycle Power Systems”, by M. Boss, the authordiscusses the exhaust end loading that is associated with STs in theprior art:

“Exhaust sizing considerations are critical for any steam turbine, butparticularly so for combined-cycle applications. There are usually noextractions from the steam turbine, since feedwater heating is generallyaccomplished within the HRSG. Generation of steam at multiple pressurelevels (intermediate pressure and/or low-pressure admissions to theturbine downstream of the throttle) increases the mass flow as the steamexpands through the turbine. Mass flow at the exhaust of a combinedcycle unit in a three-pressure system can be as much as 30% greater thanthe throttle flow. This is in direct contrast to most units with firedboilers, where exhaust flow is about 25% to 30% less than the throttlemass flow, because of extractions from the turbine for multiple stagesof feedwater heating”.

Real Estate

A combined cycle installation, although typically smaller thanconventional steam plants, still occupies a large area. The HRSGs withtheir stacks are particularly large and require a great deal of floorarea (the HRSG for one Westinghouse model 501G gas turbine isapproximately 40 feet wide, 70 feet high, and 200 feet long). With thetrend towards deregulation of electrical power, plant owners will beseeking the ideal site for their power plants. In many instances, thisis near to the electrical load, which is usually in either an urban orindustrial area. This puts the power plant close to the end user ofelectricity, and eliminates the need for high voltage transmission lines(which also require large amounts of real estate). However, availablereal estate for a large combined cycle power plant may be difficult andexpense to attain in these areas.

Some prime real estate for these combined cycle power plants will beexisting power plants that can be repowered as combined cyclefacilities. These sites have the advantage of being properly zoned withthe necessary electrical and mechanical infrastructure. The drawback isthat the site may lack the necessary real estate for a combined cyclerepowering project. Therefore, it is desirable from a space efficiencyviewpoint as well as from a cost perspective to keep plants as small aspossible.

Noise/Public Acceptance

Public acceptance is becoming increasing difficult for many utilitypower plant projects. Factors such as noise, traffic increase,unsightliness, pollution, hazardous waste concerns, and otherscontribute to public disapproval of power plants in close proximity topopulated areas. A plant that can be built smaller, quieter, with lessequipment, lower emissions, and maintain a low profile is preferred overa larger, more obvious plant. Therefore, more compact, higher “powerdensity” (power per unit volume) combined cycle power plants aredesired.

However, to meet the current trends in demand for power consumption,conventional power plants being constructed today simply replicateexisting proven plant designs to meet the increased energy consumptiondemand. No attention is currently being given to the issue of whetherplants may be redesigned to consider the ancillary issues associatedwith the public acceptance of the plants themselves.

Heat Rejection

Both conventional steam and combined cycle power plants require someform of heat rejection. This is typically to condense the low-pressuresteam from the ST exhaust back into water. This heat rejection can be tothe air, river, lake, or other “reservoir” that will absorb the heat.Since this heat rejection will have an effect on the local environmentand possibly the local biological life (i.e., fish in a river), methodsto reduce heat rejection requirements are always in demand.

Gas Turbine Performance Decay

Although combined cycle power plants demonstrate high efficiencies,these efficiencies are for “new” power plants. Since the combined cyclesin the prior art are primarily GT based, their efficiency levels arevery susceptible to GT performance decay, a phenomenon in which theefficiency of the GT degrades substantially (2% to 6%) within only ayear or two of operation. This can be a significant factor in the costof fuel as the overall combined cycle efficiency also degrades as the GTperformance decays.

OBJECTS OF THE INVENTION

Accordingly, the objects of the present invention are to circumvent thedeficiencies in the prior art and affect the following objectives:

1. Provide a combined cycle power plant that has more design flexibilitythan current offerings so that developers can have state-of-the-artfacilities, but purchase them at the capacity they need.

2. Reduce overall fuel consumption at rated output, but especially atpart load conditions, as the plant will likely spend only a smallfraction of its operating time at rated load.

3. Reduce installed cost of the power plant such that the debt serviceis substantially reduced and that financing by a bank or other lendinginstitution is much easier for the owner.

4. Leverage the time value of money with regards to capital,maintenance, and fuel costs to make the creation of power plants moreeconomically efficient and hopefully reduce the overall cost of electricpower generation.

5. Provide the ability for the power plant to meet peak demand loadswithout sacrificing normal operation efficiency or significantlyincreasing the installed cost.

6. Reduce inefficiencies and losses associated with the transmission ofpower over long distances.

7. Increase the overall reliability of the electrical grid by permittingelectrical power to be generated local to the demand during times ofpeak demand loads.

8. Reduce O&M costs. Besides fuel costs, the objective is also to reducecosts for maintenance, supplies, inventory, insurance, and otheroperating expenses.

9. Reduce the need for fuel gas compression.

10. Improve reliability.

11. Reduce air consumption and air filtering requirements.

12. Lower emissions of criteria pollutants, especially NOX.

13. Minimize the discharge of water from HRSG blowdown and othersources.

14. Utilize controls to the maximum extent feasible to increaseefficiency, reliability, and heat recovery.

15. Simplify operation and devise methods and/or strategies to increasepart load efficiency and reduce emission levels.

16. Optimize the ST efficiency by utilizing designs with improvedvolumetric efficiency and excess capacity to meet peak power demands.

17. Conserve space and land mass required to house the power plant bydesigning a compact, high power density arrangement.

18. Reduce noise, size, space requirements, and equipment to minimizethe effect the power plant has on local residents and the community.

19. Keep heat rejection to a minimum.

20. Provide for economic and space efficient retrofit of existing steampower plant and combined cycle installations so as to reduce capitalcosts and the economic burden associated with major equipment additionsand added real estate requirements.

21. Provide economic incentive for new plant construction to useenvironmentally friendly designs.

22. Design combined cycle power plants that are less susceptible to gasturbine performance decay.

These objectives are achieved by the disclosed invention that isdiscussed in the following sections.

BRIEF SUMMARY OF THE INVENTION

Briefly, the invention is a system and method permitting the use offewer and/or smaller gas turbines (GTs) and heat recovery steamgenerators (HRSGs) in a combined cycle application. This conventionalcombined cycle equipment is replaced by a larger steam turbine andcontinuously fired heat recovery steam generators to provide a varietyof economic, energy conservation, and environmental benefits.

Present technology utilizes multi-pressure HRSGs to maximize the heatrecovery from exhaust gases of a GT. This arrangement is commonly usedbecause the prior art teaches away from using continuously fired HRSGsbecause of the common belief that these configurations have lowerthermal efficiencies. Despite this commonly held belief, the presentinvention teaches that continuously fired HRSGs can be configured withthermal efficiencies on par or better than current combined cyclepractice. However, to obtain this level of efficiency, the continuouslyfired HRSGs and ST must be configured and designed differently thancurrent practice.

In several preferred embodiments of the present invention, the GTs areunchanged from the present art and exhaust to an HRSG. This HRSG,however, is designed as a single pressure level steam generator (SPLSG)(or primarily a single pressure level) which is optimized for continuousfiring to produce higher pressure steam than in conventional combinedcycle practice. In addition, the HRSG is designed to have controlledfeedwater flows through the economizer/feedwater sections to maximizeheat recovery. Also, the ST is designed as a larger unit, typical ofthat which would be found in a conventional Rankine Cycle plant, withreheat and conventional ST extraction steam fed feedwater heaters tomaximize plant thermal efficiency. This benefit of a larger ST typicalof a conventional steam plant is described by Moore in patent U.S. Pat.No. 5,649,416 which is assigned to General Electric:

“Conventional steam power plants benefit in both lower cost and higherefficiency through the economies of scale of large ratings. Atraditional rule of thumb regarding cost is that the doubling of plantrating results in a ten percent reduction in cost. The cost of one largegenerating unit according to this rule would be expected to cost on theorder of ten percent less than that for a plant with two half-sizeunits. Efficiency is also improved with increased size and powerratings. As with all turbomachinery, the internal efficiency of thesteam turbine is a strong function of the inlet volumetric flow, whichis directly proportional to the rating. Also, as is well known, thethermal efficiency of the Rankine Cycle increases with the pressure atwhich steam is generated. Increasing pressure, however, reduces thevolumetric flow of the steam at the turbine inlet, reducing the internalexpansion efficiency. The offsetting effect in overall efficiency,however, is much greater at low volumetric flow than at high volumetricflow. Therefore, an additional performance related benefit of increasingturbine size is that higher steam throttle pressure can be utilized moreeffectively.”

With the use of ample supplemental firing in the HRSG, the bottomingcycle with the present invention is given the liberty to be moreindependent from the GT operation. Therefore, the GTs can be operated atfull load while the overall plant load is modulated over a wide range ofits full load capability by only changing the supplemental firing rateand the STs load. This increases the overall plant rating when utilizinga given set of GTs, provides flexibility for the combined cycle plantrating through variation in the rate of supplemental firing, as well asincreases the overall plant thermal efficiency at part load. Inaddition, it simplifies operation, and has the potential to reduceemissions.

By designing the HRSGs to be capable of firing to 2400° F., an exemplarysingle 2-on-1 arrangement of two GTs and one large ST replaces two2-on-1 arrangements (4-on-1 arrangements are typically not availablewhen reheat is utilized due to balancing problems on the reheat lines).This exemplary configuration saves two GTs, two HRSGs, one ST, threeswitchgear, three transformers, and the accessories, real estate, andmaintenance required to support this equipment. Capital costs for thepower plant in US$/kW are thus greatly reduced using the teachings ofthe present invention.

All this is accomplished by utilizing proven turbomachinery technologyand hardware. The continuously fired HRSG with a single pressure is anovel concept for this application, but is not beyond technologicalpractice nor capability for implementation in the current art.Therefore, there are little or no compromises in reliability. Thegeneral architecture for several preferred embodiments of the presentinvention is illustrated in FIG. 13, with several exemplary embodimentshaving more detail illustrated in FIG. 9 and FIG. 15.

Improvements Over the Prior Art

The present invention solves the problems present in the prior art byachieving the following objectives:

1. Providing more design flexibility in the combined cycle power plantso that developers can still achieve state-of-the-art efficiency, butyet specify the capacity they need.

2. Reducing overall fuel consumption by improving both full load andpart load efficiency.

3. Reducing installed costs by increasing the power density of theinstallation (more power output per a given amount of equipment).

4. Reducing the overall cost of producing electricity by reducing thethree major factors associated with its production: fuel consumption,capital costs, and maintenance costs.

5. Provide temporary capacity for attaining peak loads by utilizingsupplemental firing to produce more steam, as well as having the optionto operate the ST at overpressure (inlet pressure slightly above rated)and reducing extraction steam flow to the feedwater heaters.

6. Increasing the efficiency of the power grid by permitting localgeneration of power during periods of peak loading. By permitting localpower generation during these peak periods, inefficiencies associatedwith “importing” power from other areas of a given country (and outsidea country) are reduced or eliminated. (These are energy lossesassociated with transmitting power through power transmission lines).

7. Increasing the reliability of the electrical power grid by reducingthe long haul transmission of electrical power during times of peakpower loading.

8. Reducing O&M costs, primarily by reducing the amount of equipment andsystems and utilizing equipment that has lower maintenance costs per kWhproduced (low maintenance cost STs versus high maintenance cost GTs).

9. Minimizing the need for fuel gas compression by utilizing fewer GTsand GTs with lower fuel gas pressure requirements in the cycle inconjunction with a larger ST.

10. Improving reliability by reducing the complexity of the power plantdesign.

11. Reducing air consumption by utilizing fewer GTs.

12. Lowering emissions of criteria pollutants, especially NOX, byoperating the GTs at a steady, low emissions operating point, utilizingcleaner GTs, and utilizing fewer GTs.

13. Minimizing blowdown and other discharge through higher efficiencycycles that require less steam flow per kW of electricity generated.

14. Utilizing controls to increase efficiency, reliability, and heatrecovery.

15. Simplifying operation by running the GTs at full load over a widerange of operation (total combined cycle plant output) and reducing HRSGpressure levels to only one.

16. Maximizing ST efficiency by increasing volumetric flows, especiallyin the HP and IP sections.

17. Conserving space and land mass with less equipment and higher powerdensity designs.

18. Reducing noise, size, and space requirements with less equipment.

19. Keeping heat rejection to a minimum by utilizing high efficiencycycles with less heat rejection per kWh produced.

20. Providing a combined cycle design that is more compatible withexisting steam power plants allowing for more compact and cost effectiveretrofits of these existing plants to high efficiency combined cycletechnology.

21. Minimizing air consumption, emissions of criteria pollutants, andheat rejection to the atmosphere, but providing these environmentalbenefits with lower cost than the conventional combined cycles.

22. Reducing the impact of gas turbine performance decay by utilizing acombined cycle power plant that is less dependent upon the gas turbinesand their efficiency.

BRIEF DESCRIPTION OF THE DRAWINGS

For a fuller understanding of the advantages provided by the invention,reference should be made to the following detailed description togetherwith the accompanying drawings wherein:

FIG. 1 illustrates a basic Rankine thermodynamic cycle;

FIG. 2 illustrates a schematic of a conventional prior art powergeneration system implementing the basic Rankine Cycle;

FIG. 3 illustrates a schematic of the Rankine Cycle including a reheatcycle and extraction steam feedwater heating as applied to aconventional prior art power plant application;

FIG. 4 illustrates a comparison table of efficiencies between the basicRankine Cycle and the Rankine Cycle with various efficiencyenhancements;

FIG. 5 illustrates a schematic of the basic principles of a combinedcycle;

FIG. 6 illustrates a schematic of the prior art for a combined cyclepower plant utilizing gas turbines, HRSGs, and steam turbines;

FIG. 7 illustrates a curve of heat consumption versus generator poweroutput for an industry standard General Electric (GE) Model PG7241(FA)Gas Turbine;

FIG. 8 illustrates part load performance for a General Electric combinedcycle power plant with two GE S207 GTs via graphs indicating performancecharacteristics for one and two gas turbine (GT) operation;

FIG. 9 illustrates a general arrangement of one preferred embodiment ofthe present invention as applied to the application of an electric powerplant;

FIG. 10 illustrates a tabular comparison of the efficiencies that may berealized using the teachings of the present invention as compared to theprior art;

FIG. 11 illustrates a typical graph of steam enthalpy versus temperatureat 1800 psia pressure assuming water as the motive fluid;

FIG. 12 illustrates a typical graph of gas turbine exhaust gas enthalpyversus exhaust gas temperature;

FIG. 13 illustrates a schematic of the general principles of the presentinvention as implemented in a combined cycle application;

FIG. 14 illustrates a typical graph of required log mean temperaturedifference (LMTD) versus fluid flow for a superheater and reheaterapplication;

FIG. 15 illustrates an exemplary embodiment of a combined cycle powerplant application utilizing the teachings of the present invention;

FIG. 16 illustrates an exemplary system control flowchart that may beused to control one or more heat recovery steam generators (HRSGs) asper the teachings of the present invention;

FIG. 17 illustrates an overall exemplary system control flowchart thatmay be used to provide overall power plant system control as per theteachings of the present invention;

FIG. 18 illustrates an exemplary system control flowchart which may beused to control and direct increased power plant output as per theteachings of the present invention;

FIG. 19 illustrates an exemplary system control flowchart which may beused to control and direct decreased power plant output as per theteachings of the present invention;

FIG. 20 illustrates an exemplary system control flowchart which may beused to control and direct transitional power control as per theteachings of the present invention;

FIG. 21 graphically illustrates the sources of energy inputs, losses,and efficiencies that are accounted for in an overall energy flowanalysis;

FIG. 22 illustrates a typical GE 207FA combined cycle power plantconfiguration;

FIGS. 23A and 23B illustrate tabulated performance data for a typical GE207FA 521 MW combined cycle power plant assuming a typical projectedoperation profile;

FIG. 24 illustrates a typical Westinghouse 2X1 501G 715 MW combinedcycle power plant configuration;

FIGS. 25A and 25B illustrate tabulated performance data for a typicalWestinghouse 2X1 501G 715 MW combined cycle power plant assuming atypical projected operation profile;

FIG. 26 illustrates a typical 725 MW combined cycle power plant definedby a preferred embodiment of the present invention;

FIGS. 27A and 27B illustrate tabulated performance data for a 725 MWpreferred embodiment of the present invention using a water-walled HRSGassuming a typical projected operation profile;

FIG. 28 graphically illustrates the relative part load performancedifference between a conventional combined cycle power plant and apreferred embodiment of the present invention;

FIG. 29 graphically illustrates several exemplary power plantconfigurations and their nominal range of available specified powerratings using the teachings of the present invention;

FIG. 30 graphically illustrates the basic steam cycle efficiencyrequired for an exemplary power plant configuration utilizing twoindustry standard General Electric (GE) Model PG7241(FA) Gas Turbines tomeet prior art efficiency levels over a range of power ratings;

FIG. 31A graphically illustrates a typical hourly regional system loadcurve (from “Electricity Prices in a Competitive Environment: MarginalCost Pricing of Generation Services and Financial Status of ElectricUtilities” (DOE Report number DOE/EIA-0614));

FIG. 31B graphically illustrates a typical load duration curve whichdepicts the overall long term use of rated plant capacity (data obtainedfrom Duke Energy Power Services, Inc.,http://www.panenergy.com/power/epdb2_(—)5.htm);

FIG. 32 illustrates a typical conservative weekly load profile utilizingthe data contained in FIG. 31A;

FIG. 33 graphically illustrates the part load efficiencies of severalexemplary power plants of the present invention as well as severalexamples from the prior art;

FIG. 34 tabulates an economic comparison of an exemplary power plantutilizing the teachings of the present invention to both a GE S207FAcombined cycle power plant and a Westinghouse 2X1 501G combined cyclepower plant, both from the prior art;

FIG. 35 is a typical heat balance process flow diagram for thesubcritical exemplary power plant embodiment of the present inventionused in FIGS. 26, 27A, 27B, 28, 33 and 34;

FIGS. 36, 37, and 38 tabulate some of the process data associated withFIG. 35;

FIG. 39 is a heat balance process flow diagram for theultrasupercritical exemplary power plant embodiment of the presentinvention used in FIG. 33;

FIGS. 40, 41, and 42 tabulate some of the process data associated withFIG. 39;

FIG. 43 graphically illustrates a power plant load control method thatmay be used with a combined cycle of the present invention in which twoor more GTs are utilized;

FIG. 44 tabulates data for the comparison of a retrofit of an existingsteam power plant to combined cycle technology between the preferredembodiment and the prior art;

FIG. 45 illustrates a preferred embodiment combined cycle power plantutilizing a hybrid fuel arrangement with a combustible fuel (CF) boiler;

FIG. 46 illustrates a preferred embodiment combined cycle power plantutilizing a hybrid fuel arrangement with a nuclear reactor, geothermalsteam generator, or other steam producing energy source;

FIG. 47 is an exemplary design/financing process flowchart fordetermining a preferred and/or optimal arrangement of a given inventionembodiment for a particular power plant application;

FIG. 48 is an exemplary plant economics process flowchart fordetermining a preferred and/or optimal arrangement of a given inventionembodiment for a particular power plant application;

FIG. 49 is an exemplary plant retrofit process flowchart for determininga preferred and/or optimal arrangement of a given invention embodimentfor a particular power plant retrofit application;

FIG. 50 is an exemplary hybrid fuel design process flowchart fordetermining a preferred and/or optimal arrangement of a given inventionembodiment for a particular power plant application utilizing hybridfuel;

FIG. 51 illustrates a GE three casing, four-flow steam turbine with acombined HP/IP section and two double flow LP sections.

DESCRIPTION OF THE PRESENTLY PREFERRED EXEMPLARY EMBODIMENTS

Exemplary Disclosure

While this invention is susceptible to embodiment in many differentforms, there is shown in the drawings and will herein be described invarious detailed preferred embodiments of the invention with theunderstanding that the present disclosure is to be considered as anexemplification of the principles of the invention and is not intendedto limit the broad aspect of the invention to the embodimentillustrated.

Diagrams and Flowcharts

It should be noted specifically within the context of the descriptionsgiven in this document that schematics, flowcharts, diagrams, and thelike may be augmented with components and/or steps with no reduction inthe generality of the teachings of the present invention. Similarly,components and/or steps may be removed and/or rearranged in thefollowing descriptions with no loss of generality. This notice isespecially important with respect to exemplary process flowcharts, inwhich the teachings may be used by one skilled in the computer arts togenerate control systems that are functionally equivalent, but which mayrearrange or modify the disclosed steps and processes yet achieve theresults as dictated by the present invention teachings.

Equipment

Throughout the discussion of the present invention contained throughoutthis document mention will be made to specific equipment from GeneralElectric, Westinghouse, and other manufacturers. Specifically, much ofthe disclosure makes reference to the GE model S207FA power plantcomprising GE model PG7241FA gas turbines as well as comparableequipment by Westinghouse and others. These references are exemplaryonly, and given to provide the reader who is skilled in the art aframework in which to understand the teachings of the present invention.

Rather than speak in terms of fictitious equipment which may not befamiliar to those skilled in the art, this disclosure attempts to bemore practical by illustrating the teachings of the present invention interms of equipment that one skilled in the art will be familiar with andwhich is currently in use within the electric power industry. Nothing inthis disclosure should be interpreted to limit the scope of theteachings of the present invention to a specific manufacturer or modelof equipment. On the contrary, the present disclosure should beinterpreted as broadly as possible with respect to the equipment towhich the teachings may apply.

Overview

Steam has been used for power applications for decades, dating back tosteam locomotives that burned solid fuel such as wood or coal to producepower. Up to and into the 1980's, steam power plants were stillproducing the bulk of the electrical power in the United States ineither coal, oil, or nuclear-fueled power plants.

However, by the 1980's, many smaller cogeneration power plants werebeing designed and built. These plants utilized a gas turbine as theirmain engine with a heat recovery steam generator (HRSG) connected to theexhaust of the gas turbine to recover waste heat (typically 900° F. to1200° F. exhaust gases) and convert it into steam. This steam was thenutilized for various purposes, district heating, process steam, orgeneration of additional power in a steam turbine. This plantconfiguration, gas turbine, HRSG, and steam turbine became known as acombined cycle arrangement, and due to its high efficiency, low cost,and ease of construction, has become the preferred power plant for theemerging Independent Power Producers (IPPs).

However, through evolution, this combined cycle power plant has become apower plant that utilizes the gas turbine as its prime engine and thesteam turbine as its secondary engine. An examination of the standardcombined cycle packages offered by gas turbine manufacturers today willverify this statement, as in most combined cycle plants in the priorart, the gas turbines produce about two thirds of the total poweroutput, with the steam turbines producing about the remaining one third.A review of the manufacturer's standard combined cycle offerings willillustrate this trend. The 1997 TURBOMACHINERY HANDBOOK, (USPS 871-500,ISSN 0149-4147), tabulates standard combined cycle power plantsavailable from various manufacturer's including ABB, General Electric,and Westinghouse. In most every instance, the steam turbine(s) output iswithin the range of 40% to 60% of the gas turbine(s) output. GeneralElectric informative document GER-3567G, 1996, “GE Heavy-Duty GasTurbine Performance Characteristics”, by Frank J. Brooks, provides theoutput for the gas turbines used in their combined cycle power plants.

Several preferred embodiments of the present invention recognize thecombined cycle arrangement for its high efficiency, low cost, and easeof construction. However, the present invention takes a differentperspective on the relative size of the individual engine types.Although modern gas turbines have efficiency levels in the 30 to 40%range (LHV), they require the use of an HRSG and steam turbine toachieve the combined cycle efficiency of 50 to 60% (LHV). In addition,to effectively recover the heat of the exhaust gases, these HRSGstypically have three pressure levels for the steam, high-pressure,intermediate pressure, and low pressure. The use of the intermediate andlow-pressure steam results in an overall steam cycle efficiency of only34 to 36%.

Modern large power plant steam cycle efficiencies, however, are in the45% to 50% efficiency range. To achieve these levels, the use oflow-pressure steam, as is the case with conventional combined cycles, isunacceptable. Therefore, several preferred embodiments of the presentinvention describe a method that utilizes only high-pressure steam toachieve high steam cycle efficiencies in a combined cycle configuration,yet still recovers as much heat from the exhaust gases of the gasturbine as the high efficiency, combined cycle technology in the priorart.

By this implementation, the new technology combined cycle power plantdiverges from the typical arrangement in the prior art where the gasturbine (GT) was the prime (larger) engine and the steam turbine to gasturbine power ratio was approximately 1:2, to an arrangement where thesteam turbine (ST) is typically the prime (larger) engine and the ST toGT power ratio (ST/GT) can typically be selected to be in the range of0.75:1 to 2.25:1 or greater. This ratio is easily adjusted by the designof the steam turbine, the rated amount of supplemental firing, and thesteam cycle.

During the operation of any power plant, the operations staff mustmodulate the power plant's output to the load on the system (powerconsumption by all users in the electrical grid). As the system loadfluctuates, the total power produced by all the power plants connectedto the grid must change to meet this fluctuation, otherwise, the speedof the equipment will change, and the resulting power produced will nolonger be at 60 Hz (60 cycles per second for U.S. plants, etc.). Thiswill have a dramatic effect on the equipment that the end users have inservice (i.e. electric clocks will not keep accurate time, electricmotors will not operate at appropriate speeds, etc.). Therefore, theutility and power plant personnel are responsible for maintaining aconstant frequency or speed on their equipment. To achieve this, theymust constantly change their power output to match that of the system.Note in European and various other countries this standard frequency is50 Hz, versus 60 Hz in the United States and other countries in theWestern Hemisphere such as Canada.

During the hot summer months and on extremely cold days in the winter,the system load is near its seasonal peak. Also, typically between 4 PMand 8 PM on weekdays, the system is near its daily peak. However, duringnights and weekends, the system load might only average 60% of theweekday peak. Due to these dynamics for the system load, it is uncommonfor a dispatched power plant (dispatched means controlled by the utilityto meet system load) to operate at its rated output, or any steady load,for an extended period of time. Instead, it is typically operated athigh loads during weekday peak hours (not necessarily its rated output)and at relatively low loads (approximately 60% output) for extendedhours during nights and weekends. Refer to FIGS. 31A, 31B, and 32 formore information on typical load profiles.

Therefore, to be efficient, a power plant must have the flexibility tooperate continuously at varying loads between 50% and 100%. Conventionalcombined cycle power plants are efficient, but sacrifice a great deal ofefficiency when operating at part load. This is especially true ofplants where the GT is the primary engine. In these plants, to reduceload initially from full load, the more sophisticated GTs equipped withinlet guide vanes (IGVs) will reduce airflow through the engine, thusreducing their pressure ratio. In addition, to further reduce load,these engines must reduce their turbine inlet temperatures (alsoreferred to as firing temperature) to operate at part load. Reducingthese pressures and temperatures greatly reduces the operatingefficiency of the GT engine.

To improve combined cycle plant efficiency, reduce cost, lower emissionlevels, reduce the plant real estate requirements, and simplifyoperations and maintenance (O&M), the present invention teaches the useof an HRSG optimized for continuous supplemental firing that utilizes asingle pressure level evaporator (boiler) with equal or greater ST inletpressures than are typically employed in combined cycle applications inthe prior art. In addition, it proposes the use of some features used inconventional Rankine Cycles not employed in conventional combinedcycles.

Refer to FIG. 9 for an exemplary embodiment of this new cycle. As in atypical combined cycle application in the prior art, this newarrangement utilizes one or more GTs (920) as the topping cycle powerdevice. Also, as in the typical combined cycle application in the priorart, the GT exhaust gases are fed into the HRSG. From this point,however, the cycle is changed from conventional combined cycle practice.A single pressure level HRSG is utilized rather than a multiple pressurelevel HRSG. To maximize cycle efficiency, the pressure of steam producedcan be much higher than the nominal 1800 psia typically seen. Thispressure could be supercritical (greater than 3206 psia) if desired. Forsimplicity, this discussion will focus on a sub-critical application(2400 psig rating) with an exemplary implementation example. However,performance curves for supercritical steam conditions will be includedand discussed.

Energy Flow Analysis

First, it is instructive to examine the overall energy flow in aconventional combined cycle application. From a simple energy analysis,FIG. 21 illustrates the energy flow in a combined cycle applicationwhile FIG. 10 quantifies, for the Prior Art option, the flow of energyin a conventional combined cycle plant (see the subsequent section onPreferred Embodiment Cycle Optimization for the equations used tocalculate the values in FIG. 10). This table documents performance for aGE model PG7241(FA) GT at ISO conditions with 3.0 inches H₂O inlet airpressure drop and 10.0 inches H₂O exhaust pressure drop. ISO conditionsare defined as 59° F. and 14.696 psia ambient pressure. Referring toFIG. 21, of the initial fuel input to the GT, GTI (2101) 32.31% (allpercentages based on HHV) is converted into electricity, which is the GToutput, GTO (2105). Based upon the GT exhaust gas flow and its enthalpy,only 56.21% of the input energy is sent to the HRSG, HGI (2103), meaningthat 11.48% is lost between the GT and the HRSG GTL (2102). This islikely GT generator losses, GT heat loss, gear driven accessories, motordriven accessories, windage loss, and other miscellaneous losses. Forthis example, no energy from supplemental firing will be added,therefore, SFE (2104) is zero. Of this remaining 56.21% of the GT inputenergy sent to the HRSG HGI (2103), about 10.7% (which equals(0.107)(0.5621) or 6.04% of initial GT input energy) is lost up theexhaust stack HGE (2107).

Of this remaining energy available in the GT exhaust gases to producesteam in the HRSG, 1% is considered to be lost as heat to ambient HGL(2106). Converted into terms of GT input energy, this equates to lossesof 6.04% of the GT input energy for exhaust loss and 0.50% of the GTinput energy for HRSG heat loss. This now leaves 49.67% of the GT inputenergy as energy transferred to the steam HRS (2108) which is availablefor recovery and conversion to electricity by the ST.

With a published heat rate of 6040 BTU/kWh (LHV) for a GE STAG™ S207FAplant with two GE Frame 7s and one ST, the plant efficiency based on thehigher heating value (HHV) of natural gas is 50.90%. If the GT converts32.31% of the fuel input GTI (2101) into electricity, then bysubtraction the ST must convert 18.59% of the fuel input GTI (2101)input energy into electricity. With a steam turbine generator efficiencyof 99% (or 1% loss, SGL (2110)), and a auxiliary load factor of 97.5%,and 49.67% of the fuel input HRS (2108) available to the ST cycle, thenthe basic steam cycle efficiency calculates to 38.78%((18.59/49.67)/(0.975)(0.99)). This is significantly less than the46.78% efficient operation from advanced steam cycles in a Rankine Cycleonly plant (see FIG. 4).

This steam cycle efficiency is confirmed by General Electric in theirinformative document GER-3574F (1996), entitled “GE Combined-CycleProduct Line and Performance” by David L. Chase, Leroy O. Tomlinson,Thomas L. Davidson, Raub W. Smith, and Chris E. Maslak. In discussingsupplemental firing of the HRSG, this document states

“ . . . the incremental efficiency for power production by supplementalfiring is in the 34-36% range based upon the lower heating value [(LHV)]of the fuel.”

Since supplemental firing adds heat only to the steam cycle, ittherefore follows that the steam cycle efficiency of GE's combined cycleplants is as stated.

Cycle Efficiency

The next question to be answered is how does one achieve conventionalsteam plant cycle efficiencies with the steam portion of a combinedcycle? In a review of FIG. 4, it should be observed that reheat helpsimprove steam cycle efficiency. However, reheat is already employed bymany of the high efficiency combined cycles, such as the GE STAG™ plantS207FA which utilizes two GE Frame 7s and one ST to achieve a heat rateof 6040 BTU/kWh LHV (refer to GE informative document GER-3574F (1996),entitled “GE Combined-Cycle Product Line and Performance” by David L.Chase, Leroy O. Tomlinson, Thomas L. Davidson, Raub W. Smith, and ChrisE. Maslak). Therefore, other steam cycle efficiency enhancements are theuse of higher inlet pressures, a higher volumetric flow steam turbine(for higher ST efficiency) and feedwater heating. These enhancementswill be applied in several preferred embodiments of the presentinvention.

Most modern high efficiency GTs such as the GE model PG7241(FA) haveinternal firing temperatures near 2400° F. These GTs are designed tohave exhaust gas temperatures at rated load in the range of 1100° F. Forthe PG7241(FA), at ISO conditions corrected for an HRSG exhaust loss of10.0 inches H₂O, the exhaust gas temperature is 1123° F. Thecorresponding exhaust gas flow is 3,552,000 lb/hr. At 1800 psia inletpressure, 1050° F. inlet temperature, with reheat to 1050° F.,exhausting at 1.2 inches of mercury absolute (HgA), this steam cyclewould require 1642.4 BTU/lb of heat input (reference FIG. 4, Option 2).With an inlet enthalpy of the exhaust gases of 412.6 BTU/lb, and exhaustgas enthalpy of 159.2 BTU/lb, the exhaust gases have the energy contentto produce 548,000 lb/hr of steam flow. However, closer examinationreveals a flaw in this logic. At 1800 psia, steam boils at 621° F. Sinceheat flows from higher temperatures to lower temperatures, a reasonabletemperature for the exhaust gases leaving the evaporator section wouldbe 650° F. If water preheated to an enthalpy of 648 BTU/lb was suppliedto the evaporator (an optimistic assumption), the energy required forsteam production would be 1642.4−648=994.4 BTU/lb.

However, at 650° F., the exhaust gases have an enthalpy of 281.3 BTU/lb.Therefore, the exhaust gases have the ability to boil and reheat only469,000 lb/hr (3,552,000)(412.6-281.3)/994.4. Hence, the issue becomes aheat exchange problem, as there is insufficient high level (hightemperature) energy to provide the steam at higher pressures.

Conversely, from 650° F. exhaust gas temperature to exhaust at 180° F.,there is sufficient energy to preheat 729,000 lb/hr of water((3,552,000)(281.3-159.2)/(648-53)) from the hotwell at 53 BTU/lb tosaturation enthalpy of 648 BTU/lb. Therefore, for heat recovery, in theprior art, the single pressure boiler is inefficient and either makesexcess hot water, which has little or no use in a power productionfacility, or has an HRSG exhaust temperature that greatly exceedsoptimum. This result has prompted the introduction of the multi-pressurelevel HRSG within the prior art. This arrangement makes use of theaforementioned hot water or exhaust gas energy by providing a lowerpressure evaporator section(s) in the HRSG that converts what would benon-usable hot water/exhaust gas energy to lower pressure steam.Although it has less energy content and less ability to produce power inthe ST than the high-pressure (HP) steam, this low-pressure (LP) steamnonetheless does add to the power output of the ST and serves to reducethe plant's heat consumption for a given power output (heat rate).

Supplemental Firing

Another method to alleviate the heat transfer shortcomings of a singlepressure HRSG, without adding more pressure levels as in the prior art,is to add energy at or near the inlet of the HRSG through supplementalfiring. However, the current teachings are that supplemental firingreduces overall plant thermal efficiency. This is noted by Moore in U.S.Pat. No. 5,649,416 in which he states

“Supplemental firing of the heat recovery steam generator can increasetotal power output and the portion of the total power produced by thesteam turbine, but only with a reduction in overall plant thermalefficiency.”

In addition, Rice, in U.S. Pat. No. 5,628,183 states

“Supplementary firing in front of the HRSG does not offer a viablesolution towards higher cycle efficiency.”

Thus, the prior art specifically teaches away from this technique ofsupplemental firing. In addition, Rice references other documents by GEand Westinghouse that concur with his statement. GE informativeliterature, GER-3574F (1996), entitled “GE Combined-Cycle Product Lineand Performance” by David L. Chase, Leroy O. Tomlinson, Thomas L.Davidson, Raub W. Smith, and Chris E. Maslak states

“ . . . the incremental efficiency for power production by supplementalfiring is in the 34-36% range based upon the lower heating value of thefuel.”

This states that although combined cycle efficiency is 56% based on thelower heating value (LHV) of the fuel at full load, power producedthrough supplemental firing is added at an efficiency equal to or lessthan 36% LHV.

Also in this document, (GE informative document GER-3574F, 1996,entitled “GE Combined-Cycle Product Line and Performance” by David L.Chase, Leroy O. Tomlinson, Thomas L. Davidson, Raub W. Smith, and ChrisE. Maslak), another source which identifies supplemental firing as adetriment to efficiency (heat rate) is Table 14 which indicates thatHRSG supplemental firing can increase combined cycle plant output in theprior art by 28%, but only with an increase in overall combined cycleheat rate (specific fuel consumption) of 9%.

Present Invention Energy Flow

It will now be instructive to reexamine the overall energy flow in acombined cycle application as utilized in several embodiments of thepresent invention. From a simple energy analysis, FIG. 21 illustratesthe energy flow in a combined cycle application while FIG. 10 quantifiesthe energy flow in a preferred exemplary embodiment combined cycle (seethe following section on Preferred Embodiment Cycle Optimization for theequations used to calculate the values in FIG. 10). Again, GTperformance is for a GE model PG7241(FA) GT at ISO conditions and 3.0inches H₂O inlet air pressure drop and 10.0 inches H₂O exhaust pressuredrop. Referring to FIG. 21, of the initial fuel input to the GT, GTI(2101), 32.31% (all percentages based on HHV) is converted intoelectricity, which is the GT output, GTO (2105). Based upon the GTexhaust flow and enthalpy, only 56.21% of the GT input energy is sent tothe HRSG HGI (2103), meaning that 11.548% is lost between the GT and theHRSG, GTL (2102). Of this remaining 56.21% of the GT input energy sentto the HRSG, about 10.7% of it is lost up the exhaust stack HGE (2107),leaving 50.17% of GT input energy to the HRSG. To this point, the energyflow is unchanged from the prior art.

To ensure maximum heat recovery in the HRSG, several of the preferredembodiments of the present invention prescribe increasing the feedwaterflow through the HRSG until there is a sufficient balance of heat gainby the feedwater to match the necessary heat loss from the exhaust gasesfor optimum heat recovery (i.e. reduce HRSG exhaust gas temperature toapproximately 180° F.). Secondly, through the addition of fuel at theHRSG inlet (supplemental firing), the exhaust gas energy in the HRSG israised until there is sufficient energy to convert most or all thefeedwater flow into steam. Using Option 3 from FIG. 4, heat must beadded at 1633.9 BTU/lb to produce the desired steam conditions. Sincethe heat capacity of the exhaust gases is approximately 0.25 BTU/lb/°F., and the heat capacity of the returning condensate is approximately1.0, the steam flow should be near 0.25 lb of steam per lb of exhaustgas flow. For two GE frame 7 GTs this yields a steam flow of 1,776,000lb/hr.

To produce this amount of steam will require 2894 MMBTU/hr (millionBTU/hr). With 1% loss to ambient from the HRSG, HGL (2106), the heatinput requirement becomes 2923 MMBTU/hr. With exhaust loss, thenecessary HRSG input energy required to produce this steam is 87.99% ofthe GT input energy. Since the HRSG input energy HGI, (2103) minus theHRSG exhaust loss, HGE (2107), equals 50.17% (56.21(−6.04), of GTI(2101), an additional amount of energy equal to 31.78% of the GT inputenergy must be added as heat from supplemental firing SFE (2104),yielding a total of 81.95% of GTI (2101). Adjusting for a 1% loss toambient, HGL (2106), 81.13% of GTI (2101), the GT input energy, convertsto steam. This steam is now available for conversion to electricity bythe ST.

With a ST for use in a preferred exemplary embodiment, higher pressure,reheat, and feedwater heating may all be employed. In addition, the STrating will be an estimated 2.5 times that of a conventional combinedcycle plant. This would lead a reasonable designer to use the steamcycle efficiency of 44.39% as shown for Option 5 in FIG. 4 (as perMoore, ST efficiencies increase with rating, but for demonstrationpurposes, an overall 90% has been retained for this example).

Utilizing a 44.39% efficient basic steam cycle, 36.01% of the availableheat is converted to shaft horsepower, utilizing a factor of 97.5% toaccount for auxiliary loads and a 99% efficient generator. ST electricaloutput is therefore 34.76% of GTI (2101), GT input energy. With the GToutput, GTO (2105), equal to 32.31% of GTI (2101), the ST output equalto 34.76% of GTI (2101), and with an additional supplemental fuel inputof 31.18% of GTI (2101), combined cycle efficiency therefore becomes(output divided by input) ((0.3231+0.3476)/(1+0.3118)) which equals50.90%. Utilizing only two (2) FWHs in the cycle, the efficiency of thisexemplary preferred embodiment is on par with the GE conventionalcombined cycle plant. For supercritical applications, the overallcombined cycle efficiency in several of the preferred embodimentsincreases to 51.75% and lowers the heat rate to 5942 BTU/kWh LHV(reference FIGS. 10 and 21).

Therefore, from an overall energy perspective, it is apparent thatsupplemental firing is NOT detrimental to overall combined cycleefficiency IF a commensurate increase in bottoming cycle efficiencyaccompanies the supplemental energy addition to the bottoming cycle.

Preferred Embodiment Cycle Optimization

As stated, one of the major improvements for several of the preferredembodiments of the present invention is the flexibility. Withsupplemental firing, the new combined cycle power plant can be designedwith a combination of various gas turbines together with a customdesigned steam turbine(s) to provide a much wider range of applicationfor the new combined cycle power plant.

Since efficiency is defined as output divided by input, the energy flowanalysis can be used to determine the steam cycle efficiency required ata given rating. Therefore, for overall combined cycle efficiency, theoutput is the combination of both the steam turbine and gas turbine(s)electrical output. The input is the total of the GT input energy alongwith the energy added to the duct burners through supplemental firing.Therefore, referring to FIG. 21, the equation for combined cycleefficiency (.eta.) of several of the preferred embodiments of thepresent invention is given by the relation: $\begin{matrix}{\eta = \frac{\left( {{GTO} + {STO}} \right)}{\left( {{GTI} + {SFE}} \right)}} & (2)\end{matrix}$

where

GTO=gas turbine(s) electrical output

STO=steam cycle electrical output

GTI=gas turbine(s) input energy

SFE=HRSG input energy through supplemental firing.

In the above exemplary equation, the values of GTO, GTI, and SFE aretypically known. The unknown variable is the steam cycle electricaloutput, STO. This number is a function of several other inputs,including steam turbine generator efficiency, HRSG exhaust loss,auxiliary load factor, and finally steam cycle efficiency. First, it isnecessary to calculate the amount of energy that is transferred to thesteam from the HRSG. This is defined as HRS (2108) and is calculatedfrom the following equation:

HRS=[HGI+SFE−HGE] (1−HGL)  (3)

where

HRS=HRSG energy transferred to steam

HGI=GT exhaust heat

SFE=supplemental firing heat

HGE=HRSG exhaust loss

HGL=heat loss to ambient

The above exemplary equation essentially calculates the heat into thesteam as the sum of: the GT exhaust heat, plus the heat added fromsupplemental firing, minus the HRSG exhaust loss, with a correction forheat loss to ambient. This now defines the quantity of energy availableto the steam cycle. To determine the electrical output from this energy,STO (2111), this energy input must be adjusted for the steam cycleefficiency, SCE (2109), the ST generator losses, SGL (2110), and theauxiliary loads, AXF (2112). The equation for steam turbine generatoroutput becomes:

STO=HRS×SCE×AXF×(1−SGL)

where

HRS=HRSG energy transferred to steam

SCE=steam cycle basic efficiency

AXF=auxiliary load factor

SGL=steam turbine generator losses=(1−steam generator efficiency (SGE))

The steam cycle efficiency value therefore converts steam energy into STshaft power, which is then corrected to steam cycle electrical output bycorrections for both the generator efficiency and the reduction of poweroutput by the auxiliary loads.

Knowing these equations, and also knowing the desired output for a givenGT arrangement (see FIG. 29 for range of outputs of several of thepreferred embodiments of the present invention), the required steamcycle efficiency can be determined which will yield a preferredembodiment combined cycle plant efficiency equal to that of theconventional (lower rating) combined cycle plant from the prior artwhich was based on the same GTs. FIG. 30 illustrates the steam cycleefficiencies that are required as the combined cycle plant described byseveral of the preferred embodiments of the present invention isincreased in rating. Note that the parameter along the horizontal axisis the ratio of ST power output to the total of all GT(s) power output.

Utilizing FIG. 29, FIG. 30, and the aforementioned equations for steamcycle efficiency and overall plant efficiency, a design engineer skilledin the art can determine which GT combination is most favorable fromboth an energy efficiency and economic perspective, and determine therelative complexity of the steam cycle (subcritical steam conditions,amount of feedwater heating, inlet temperatures, etc.) that will yieldthe desired overall plant efficiency. Refer to FIGS. 47-50 illustratinga plant design/construction method for the selection, design, andfinancing of the preferred embodiment of the present invention.

Preferred Embodiment Flexibility

As previously mentioned, flexibility is one of the major advantages tothe present invention. From an examination of FIG. 30, it can be seenthat at lower ST/GT ratios, a steam cycle of more moderate efficiencycan be utilized to provide on par plant efficiency utilizing theteachings of the present invention. However, it would be possible at lowST/GT ratios to utilize ultrasupercritical steam conditions to exceedthe efficiency of a combined cycle power plant from the prior art. Ifthe exemplary preferred embodiment in FIG. 26 at 725 MW were to use anultrasupercritical bottoming cycle, the heat rate would be reduced from6006 BTU/kWh to 5912 BTU/kWh.

However, unlike preferred embodiments with higher ST/GT ratios, thisconfiguration yields less operational flexibility than preferredembodiments with higher ST/GT ratios. With these lower ratios, thecontrol of the preferred embodiment will be more like that of the priorart in that the GTs will need to be modulated to control plant load at ahigher plant operating point. Depending upon the economics, highefficiency, low efficiency, or capital costs will determine which ST/GTratio is ultimately chosen by the power plant developer.

Preferred Embodiment Potential Ratings and ST-GT Ratio

FIG. 30 illustrates the approximate steam turbine rating increases thatare attainable from several of the preferred embodiments of the presentinvention. A conventional combined cycle power plant from the prior artcould have a ST output that is nominally 55% of the total GTs output.Therefore, total plant output could be defined as 1.55 (1.00 for GTsplus 0.55 for the ST) of GTs output. With this example of several of thepreferred embodiments of the present invention, the ST could be designedto be as much as 2.1 times the output of the GTs, such that total plantoutput is 3.1 (1.0 for GTs plus 2.1 for the ST) times the output of theGTs.

This example of a preferred embodiment of the present invention has arating that is 3.1/1.55=2.0 times that of the prior art. To remain onpar in efficiency with the prior art, however, the basic steam cycleefficiency needs to be 48.3% (refer to FIG. 30). With supercriticalsteam conditions, advanced steam parameters, and feedwater heating,basic steam cycle efficiencies can come close to this benchmark.Therefore, several of the preferred embodiments of the present inventionhave the ability with certain gas turbine arrangements to nearly doublecombined cycle power plant output as compared to the prior art,drastically reduce the amount of hardware that would have been requiredin the prior art to attain this output, yet still manage to achieveefficiency levels that are on par with the prior art.

Since the present invention teaches the use of a single pressure levelHRSG, and to efficiently utilize a single pressure level HRSG, thefeedwater flow through the low temperature section of the HRSG must beadequate to absorb the GT exhaust gas energy, analysis has shown that anST/GT ratio minimum of approximately 0.75 is required to meet thisobjective. Assuming a relative GT power output of 1.0 and a ST/GT powerratio of 0.75, yields a total plant power output of 1.75 times the GToutput, resulting in a GT to total power output of (1.0/1.75) orapproximately 0.57 or 57% of the total plant power output.

Design Limitations

Although several of the preferred embodiments of the present inventionoffers a more expansive range of combined cycle ratings for a given setof gas turbines than was available in the prior art, there are stilllimitations on the design of these new technology combined cycle powerplants. Some of these limitations are as follows:

1. Above approximately a 1600° F. duct-fired gas temperature at itsinlet, the HRSG will require a more expensive water-wall construction.

2. With water-wall construction, the HRSG may be limited toapproximately a 2400° F. duct-fired gas temperature.

3. The HRSG exhaust gases must contain sufficient oxygen to support thecombustion of additional fuel.

4. The duct burners that provide additional heat input to the HRSG mustbe able to maintain low NOX levels even at high prescribed firing rates.

5. The cycle must be designed to operate within the steam turbine designparameters for pressure and temperature.

6. The cycle must be designed so as to maintain the proper efficiency,cost, emissions, or other limiting parameters that may exist to make theproject economically and environmentally acceptable.

Considering these limitations, FIG. 29 illustrates an approximate rangeof rated power for combined cycle power plants described by several ofthe preferred embodiments of the present invention. Note that thesepower plants are based upon either one or two GTs and cover a range fromless than 150 MW up to 1050 MW. FIG. 29 is not meant to represent allpossible GT combinations which can utilize the preferred embodiment ofthe present invention, but represents only a sample of various GTs fordemonstration purposes.

Impact of Economic Considerations on Plant Design

All power plant design engineers skilled in the art review numerousplant design options for their relative economic merit before selectinga final configuration for a power plant. This is true with the combinedcycle plants from the prior art and will be true of combined cycleplants utilizing the system and method described by several of thepreferred embodiments of the present invention. New plants must becommercially feasible if they are to be constructed.

The power plant design engineer may examine alternatives such as a lowcost cell type cooling tower with a high auxiliary load (electric motordriven fans) versus a high cost hyperbolic style cooling tower with onlya small auxiliary load (natural draft air flow, no fans required). Thisbecomes an economic evaluation of the energy saved versus the capitalcost expended to save said energy. Based upon current and projectedeconomic factors for energy costs, capital costs, and other factors, thedeveloper of the power plant project will select the most economicalarrangement. The most efficient selection from an energy conservationperspective is not always the most economical selection.

These same type of evaluations will need to be presented with several ofthe preferred embodiments of the present invention. Althoughultrasupercritical steam conditions yield higher steam cycleefficiencies, the incremental savings in fuel may not outweigh the addedcost for the more intricate hardware. If interest rates are high,several of the preferred embodiments of the present invention will allowlarge capacity increases with only a nominal percentage increase inprice. With low fuel costs, larger plants without the commensurateincrease in steam cycle efficiency may be appropriate. Again, several ofthe preferred embodiments of the present invention allow the designengineer skilled in the art along with the plant developer to chose froma wider range of alternatives to find the most commercially viableoption for the power plant.

With several of the preferred embodiments of the present inventionbecoming primarily a steam plant rather than primarily a GT plant, thereare a couple economic evaluations that are usually of key interest.Since these steam turbines will be large and have high exhaust endflows, they typically utilize either one, two, or three exhaust casings,each of which has a double flow arrangement. FIG. 51 is an illustrationof a General Electric (GE) steam turbine a from GE informative documententitled “Steam Turbines for Large Power Applications” by John K.Reinker and Paul B. Mason (General Electric Reference GER-3646D, 1996).The casing to the left is the combined HP/IP section, while the twolarger sections to the right are the double flow exhaust (LP) sections.Differing sizes of exhaust casings are available which are designedaround the blade lengths in the last stage. There can be substantialcost differentials between the different exhaust casings.

The selection of the steam turbine last stage blade height, exhaustcasing size, and number of exhaust casings is one very common economicevaluation for a large steam plant. The steam cycle may become moreefficient by an increase to the next larger exhaust casing or perhapseven through the addition of another exhaust casing. However, theincremental increase in steam cycle efficiency must be weighed againstthe increase in cost for the additional hardware. Another factor thatcomes into play is the sizing of the condenser and heat rejectionequipment. Again, lower exhaust pressures yield higher steam cycleefficiencies, but the cost of the equipment to provide incrementalreductions in exhaust pressure must not outweigh the fuel savings.

In consideration of the economics of operation, the developer mustprovide the design engineer with an operation scenario for the new powerplant. Since the system electrical load is very dynamic and constantlychanging, a load profile needs to be established which exemplifies theload on the plant as a function of time. FIG. 31A is from a U.S.Department of Energy report numbered DOE/EIA-0614 entitled “ElectricityPrices in a Competitive Environment: Marginal Cost Pricing of GenerationServices and Financial Status of Electric Utilities”. FIG. 31Aillustrates a typical load profile for a system (electrical grid) on anhourly basis for a single day. On a weekly basis, this profile wouldindicate lower load on weekends and holidays, and on an annual basis,there would be adjustments for seasonal changes. Since most power plantswill operate a majority of their lifetime at partial load, the optimumeconomical arrangement results from designing these plants to be mostefficient at some average or mean load point of operation, versus at theplant's rating.

This is noted by M. Boss in GE informative document GER-3582E (1996),entitled “Steam Turbines for STAG™ Combined-Cycle Power Systems”. Inthis paper, the author explains that although the efficiency of thesteam cycle may be maximized when the ST exhaust annulus velocity at thelast stage blade is approximately 550 feet per second, the economicoptimum is typically with an exhaust annulus velocity of between 700 and1000 feet per second at the rated point of the ST. James S. Wright, inGE informative document GER-3642E (1996), entitled “Steam Turbine CycleOptimization, Evaluation, and Performance Testing Considerations”provides an evaluation for steam turbine exhaust casing selection. Inthe example, the selection is made between three different sized exhaustcasings, with the efficiency of the exhaust casings increasing with eachlarger size. The largest casing is not selected because its incrementalgain in efficiency does justify its added cost per the economicparameters. By the same token, the smallest casing is not selectedbecause its incremental savings in capital cost does not justify thelarge loss in efficiency. Therefore, the medium sized casing is selectedbecause it is the economic optimum.

Single Pressure HRSG

To make an HRSG effective at a single pressure level, its designeffectiveness must first be examined. FIG. 11 is a curve of steamenthalpy versus temperature for a pressure of 1800 psia. As can bereadily seen, the heat content of the steam is not a linear functionwith respect to temperature. This phenomenon greatly complicates theheat transfer with the exhaust gases that have a nearly linearcharacteristic (see FIG. 12). As is seen in FIG. 11, at the boilingpoint of 621° F., the water/steam mixture increases in enthalpy from 648BTU/lb to 1154 BTU/lb without any increase in temperature. The heatabsorbed in this section of the HRSG (evaporator) will be much greaterthan any other section for a given temperature change.

Between the temperatures of 100° F. and 400° F., the average heatcapacity of water is 1.014 BTU/lb/° F. This value is essentially linearand changes only slightly with pressure. Therefore, heat transfer inthis region between the water and exhaust gases will be relativelyconsistent.

To maximize the effectiveness of heat recovery in the HRSG, and to stillprovide the maximum amount of steam to the ST, a system control methodis required that optimizes the feedwater/steam flow through each sectionof the HRSG. This optimization scheme will be programmed in the powerplant's DCS control system.

System Control

There are numerous possible control techniques for the ST, however, twopopular methods are flow control and sliding pressure control. With flowcontrol, the ST includes a set of valves that is controlled to maintaindesign inlet pressure. With sliding pressure operation, the inletpressure to the ST is allowed to “slide” or change with the load (steamflow) to the ST. For combined cycle plants, where heat recovery isemployed, it is often advantageous to use sliding pressure control. Thiscontrol method allows for high volumetric flows in the steam turbine byutilizing lower specific volume steam (lower pressure) at part load.This maintains ST efficiencies at or near design levels. In addition,lower pressure steam boils at a lower temperature than higher pressuresteam, therefore, the lower temperature exhaust gases in the HRSGassociated with lower loads can produce more steam.

Energy Utilization

As demonstrated previously, in order to produce high-pressure steam inthe HRSG, it is necessary not only to have the overall energy content toproduce the steam (total required BTUs), but the energy must be at theappropriate temperature to affect the necessary heat transfer. Inaddition, it is desired to maximize the use of waste heat, and notproduce large quantities of hot water or greatly increase HRSG exhausttemperature above its optimum point. The use of supplemental firingbecomes extremely useful in meeting these goals.

It is now useful to consider the concept of a single pressure HRSG usedwith a GT. As previously demonstrated, this arrangement, when designedfor an HRSG exhaust temperature of 180° F., would produce either anexcess of hot water or a high HRSG exhaust gas temperature in the priorart. Again, this is due to the non-availability of sufficient heat inputat the higher temperatures, and an overabundance of heat available atthe lower temperatures. For illustration purposes, consider an HRSG thatadded heat to the exhaust gases without an increase in temperature (nota likely arrangement for a GT/HRSG assembly). Imagine that to add heat,the HRSG was designed to ingest more fuel and more air, but without anincrease in its inlet temperature. This scenario would provide for theproduction of a larger quantity of GT exhaust gases and thus a largerquantity of steam. However, it also would provide for a larger quantityof hot water. Effectiveness of the HRSG would not be changed, only itscapacity would be increased proportional to the heat addition.

This concept is important, because not only does it apply to the HRSG,but it applies to the conventional combined cycle practice whensupplemental firing is utilized. In the prior art, supplemental firingincreased steam flows, but did not improve the effectiveness of thesteam cycle.

Due to excess oxygen in the exhaust gases from a GT (oxygen levelsreduce from 21% O₂ in ambient air to approximately 12-15% at the GTexhaust at full load), fuel can be burned directly in the HRSG withoutthe need for additional air. This practice allows supplemental firing toincrease the temperature of the exhaust gases. The combustion and heatrecovery process for supplemental firing is essentially 99% efficient,as only 1% of the HRSG heat input is lost to ambient surroundings. Thisis a dramatic improvement over conventional Rankine cycle boilers thatmight only be 80 to 90% efficient. The primary reason for this largedifferential in efficiency between the conventional Rankine cycleboilers and HRSGs is that conventional Rankine cycle boilers ingest coldambient air for combustion and may then exhaust in the range of 350° F.to 400° F., versus the HRSG which receives preheated GT exhaust gases attemperatures between 800° F. and 1200° F., and then exhaust in the rangeof 160° F. to 200° F.

This increase in the energy level of the exhaust gases as a result ofsupplemental firing, greatly improves the ability (heat transfercapability) of these exhaust gases to produce high-pressure, hightemperature steam. In addition, more energy at the high end of the HRSGoffsets or balances the excess energy at the low end of the HRSG typicalin the combined cycle from the prior art.

In other words, additional heat input at the HRSG inlet that increasesthe exhaust gas temperature, can be transferred to the feedwater flowthat had insufficient energy to become HP steam. Not only is the overallsteam flow increased, but the effectiveness of the steam cycle is alsoincreased by producing a higher proportion of HP steam. Thus, theaddition of fuel into the bottoming cycle, as well as providingadditional heat input, can be used to increase the overall effectivenessof the bottoming cycle.

System Overview

FIG. 13 is a conceptual schematic for a combined cycle application withheat addition to the bottoming cycle. In FIG. 13, the topping cyclefluid (TCF) (1301) enters the topping cycle engine, (TCE) (1302) wherefuel and/or heat (CFT) (1303) is added to raise its temperature. Thefluid performs work that is converted by the topping cycle engine intoshaft horsepower. This shaft horsepower drives the topping cycle load,(TCL) (1304). This load could be an electrical generator, pump,compressor, or other device that requires shaft horsepower. Theexhausted fluid from the topping cycle engine is directed through andexhaust line (1305) to a heat recovery device (HRD) (1306). In addition,fuel and/or heat (CFB) (1314) is added to the topping cycle fluid at thepoint where it enters the heat recovery device. After passing throughthe HRD, the topping cycle fluid exhausts to an open reservoir (1307).

For this example, the topping cycle is an open cycle. In other words,the topping cycle fluid is taken from a large reservoir and dischargesto that same reservoir. The heat recovery device (1306) captures aportion of the topping cycle exhaust energy and transfers it to thebottoming cycle fluid (BCF) (1308). In this example, the bottoming cyclefluid is heated at a single pressure level, a high-pressure (HP) line(1309). This fluid then travels to the bottoming cycle engine (BCE)(1310) where it produces shaft horsepower to drive the bottoming cycleload (BCL) (1311). Again, this load could be an electrical generator,pump, compressor, or other device that requires shaft horsepower.

From the bottoming cycle engine, the bottoming cycle fluid enters a heatexchanger (HEX) (1312) where heat is rejected. The bottoming cycle fluidthen enters a fluid transfer device (FTD) (1313) where it is thenreturned to the heat recovery device. For this example, the bottomingcycle is a closed cycle, meaning that the bottoming cycle fluid iscontinuously circulated within a closed loop.

The present invention exemplary embodiment illustrated in FIG. 13contrasts to FIG. 5 in two major ways:

1. Fuel and/or heat is added (1314) to the heat recovery device (1306)which is not added in FIG. 5; and

2. In FIG. 13 there is only one fluid, HP fluid (1309), which issupplied to the bottoming cycle engine (1312), versus HP fluid (509), IPfluid (510), and LP fluid (511) in FIG. 5 which are supplied to thebottoming cycle engine (512).

By utilizing fuel and/or heat addition to the bottoming cycle, not onlyhas the energy to the bottoming cycle increased, but so has the cycle'seffectiveness, as now all the IP and LP fluid has been upgraded to HPfluid. This HP fluid has the ability to do more work per unit mass flowthan either the IP or LP fluids.

There are a number of different fluids that could be applied to theconceptual combined cycle arrangement, including water, air, steam,ammonia, refrigerants, mixtures, and many others. The intent of apreferred exemplary embodiment is not to limit the number of cycles usedin the combined cycle, limit the fluids in the combined cycle to anyspecific fluid, limit the fluid pressures that may be utilized, or limitany cycle to being either an open or closed cycle, but to demonstratethat the process of upgrading thermal efficiencies of combined cyclescan often be accomplished through the strategic use of additional fueland/or heat input.

Heat Transfer Analysis

HRSG LP Economizer Section

As mentioned previously, the problem in producing HP steam inconventional combined cycle power plants is the distribution of theenergy between the exhaust gases and the steam being produced. Inaddition, to optimize heat recovery, it is desired to have the exhaustgas temperature at the HRSG exit to be optimum. Therefore, a morein-depth look at the heat recovery process must be made.

To optimize heat recovery in the lower temperature regions of the HRSG(approximately 470° F. exhaust gas entering temperature to the 180° F.exhaust gas exit temperature range), a sufficient amount of heat must beremoved by the pressurized feedwater. The average heat capacity of theexhaust gases in this range (470° F. to 180° F.) is 0.257 BTU/lb/° F.(note that this value can vary slightly with exhaust gas oxygencontent/amount of supplemental firing). Between the temperatures of 100°F. and 400° F., the average heat capacity of water is 1.014 BTU/lb/° F.Therefore, to obtain an increase in feedwater temperature to correspondto a commensurate decrease in exhaust gas temperature, the flow ratioshould be (1.014/0.257) or 3.95 lbs of exhaust gas per lb of feedwaterin this temperature range of the HRSG. A flow ratio at or near thisnumber will optimize heat recovery for this section of the HRSG. Changesin parameters such as exhaust gas oxygen content, inlet watertemperature, and other factors can be monitored in the plant DCS controlsystem and the optimum feedwater flow through each section of the HRSGcan be calculated and controlled.

Experience has determined that providing cold water temperatures at theinlet to the LP economizer section (feedwater directly from thecondenser) can have detrimental effects on the life of the economizercomponents. This is due to corrosion problems in the economizer as aresult of tubes and fins in the economizer being colder than the dewpoint of the exhaust gases of the HRSG. Since these components aretypically constructed of a carbon steel or low alloy steel, thecondensed moisture on the tube and fin surfaces corrodes away thesecomponents and reduces heat exchanger effectiveness. Two common methodsare utilized to alleviate this problem. One is to use a feedwaterpreheater to introduce warmer water into the economizer. The othermethod is to construct a portion of the LP economizer from non-corrodingmaterial, such as stainless steel. Either method is acceptable, and theone selected is usually the one that is determined to be economicallyoptimum.

HRSG HP Economizer Section

The HP economizer section of the HRSG heats the feedwater fromapproximately 400° F. (exit of the LP economizer), ideally to thesaturation temperature of the pressure in the evaporator section. Usingan average pressure for this example of 1800 psia, the saturationtemperature at this point is 621° F. In this range, the average heatcapacity of the feedwater is 1.230 BTU/lb/° F. To heat this water, GTexhaust gases will need to enter the section approximately 50° F. abovethe feedwater exit temperature, or 671° F. The average heat capacity ofthe exhaust gases in this range (671° F. to 470° F.) is 0.264 BTU/lb/°F. (note that this value can vary slightly with exhaust gas oxygencontent/amount of supplemental firing). Therefore, for this section ofthe HRSG, the flow ratio should be (1.230/0.264) or 4.66 lbs of exhaustgas per lb of feedwater. Since this flow ratio does not match with theLP economizer optimum flow ratio, an adjustment will need to be made tocompensate for this mismatch (differing optimum flows through eachsection).

HRSG Evaporator Section

The evaporator section (sub-critical applications) is unique from othersections in the HRSG in that its inlet and outlet temperatures areessentially constant (for constant pressure operation). This addsstability to the heat exchange process, and the Log Mean TemperatureDifference (LMTD) fluctuates less with variations in flow than that ofother sections since the outlet temperature is essentially constant. TheLMTD is a non-linear heat transfer variable that is used to determinethe heat transfer capability of a heat exchanger.

Due to this constant temperature factor, the sections downstream of theevaporator, the HP and LP economizers, see relatively constant (slightvariation with pressure/load) input temperatures. However, in severalpreferred embodiments, supplemental firing will greatly alter the inlettemperatures to the evaporator section, as well as the superheater andreheater sections. These increasing and decreasing temperatures willdetermine the steam flow through the HRSG, and ultimately, the SToutput. Therefore, unlike the economizer sections, an optimized flowratio is not truly applicable for the upstream sections of the HRSG.

Since the evaporator section of the HRSG absorbs a major share of theheat available, and actually produces the steam, its output is modulatedmostly by the section exhaust gas inlet temperature, which is largely afunction of the HRSG exhaust gas inlet temperature. Therefore, thecontrol of this section is done primarily through fuel input.

HRSG Superheater and Reheater Sections

These sections are similar in that they both heat steam to a highertemperature. The superheater section receives saturated steam from theevaporator section and heats it to the HP turbine inlet temperature. Adesuperheater is used at the exit of this section to control thetemperature to the desired value.

The reheater section receives steam from the HP turbine section exhaustand reheats it back to the IP turbine inlet temperature. A desuperheatercan be used at the exit of this section to control the reheattemperature, but does so at a cost in cycle efficiency. This is noted byEugene A. Avallone and Theodore Baumeister III in MARKS' STANDARDHANDBOOK FOR MECHANICAL ENGINEERS (NINTH EDITION) (ISBN 0-07-004127-X,1987) in Section 9-24 through 9-25 which states:

“The attemperation of superheated steam by direct-contact water spray .. . results in an equivalent increase in high-pressure steam generationwithout thermal loss . . . Usually, spray attemperators are not used forthe control of reheat-steam temperature since their use reduces theoverall thermal-cycle efficiency. They are, however, often installed forthe emergency control of reheat steam temperatures.”

FIG. 14 is a set of curves illustrating the heat requirements for thesuperheater and reheater sections as a function of flow. These curves donot include small effects for desuperheating, extraction flows, heatloss in the pipe, or other minor adjustments. Notice that both thesesections require proportional amounts of heat with flow (ST load)changes. Therefore, it may be advantageous, although not necessary, tobuild these two sections as one in the HRSG, each with its ownappropriate heat exchange surface area.

HRSG Surface Areas

In order to obtain the necessary heat transfer from the GT exhaust gasesto the water/steam, it is required that sufficient amounts of heatexchange surface area be provided in each section. The controllingequation that describes this overall heat exchange is

Q=U×A×LMTD

where

Q=heat transferred in BTU/hr

U=overall heat transfer coefficient in BTU/hr/ft²/° F.

A=total surface area in ft²

LMTD=log mean temperature difference

with the log mean temperature difference (LMTD) being defined as

LMTD=(GTTD−LTTD)/ln(GTTD/LTTD)

where

GTTD=greater terminal temperature difference

LTTD=lesser terminal temperature difference

The terminal temperature differences are

1. the temperature of the exhaust gas into an HRSG section minus thewater or steam temperature out, and

2. the temperature of the exhaust gas exiting an HRSG section minus thewater or steam temperature in.

Obviously, the larger value is the GTTD and the smaller value is theLTTD. If they are equal, then either one equals the LMTD. If either theGTTD or the LTTD become too small, the surface area, A, must become verylarge to compensate. Since the surface area is essentially the totaleffective surface area of all the tubes and fins in the HRSG section,adding area adds size, weight, and cost to the HRSG.

The other factor in the heat exchange equation, U, is based upon thesurface coefficient of heat transfer between the water/steam and thetube inner wall, the heat conductance of the tube material and itsthickness, and the surface coefficient of heat transfer between theexhaust gases and the tube outer wall.

For general purposes, the controlling factor in this equation is thesurface coefficient between the exhaust gases and the tube outer wall.This is because it is the largest resistance to heat transfer, and likea group of resistors in series in an electrical circuit, the largestresistance controls the flow. Therefore, factors that have the greatesteffect in changing the outer heat transfer coefficient are of the mostconcern to engineers designing the HRSG and selecting the areas for eachsection.

From a control standpoint, selection of the areas in each section iscritical, because once the HRSG is built, these areas cannot be changed,but become a fixed value. Factors which affect changes in the value of Uare those which change the velocity of the exhaust gases over the tubesurfaces. The predominant deviation is a change in the exhaust gas flow.Since the GT is a constant volume machine, this occurs with changes inthe ambient air temperature. In addition, it occurs with load changes onthe GT. If these factors can be minimized, the HRSG can be more readilydesigned for operation within a narrow band and better optimized.

As will be illustrated in the example of a preferred exemplaryembodiment of the present invention, the disclosed system and methodallows for GT operation at full load (temperature control) over a widerange of total combined cycle plant load. This contrasts sharply to theprior art that utilized changes in GTs load to modulate the overallcombined cycle plant load. Therefore, at most operating points, the onlysignificant changes in HRSG flow will be attributed to ambienttemperature changes (fuel from supplemental firing adds less than 1% tothe exhaust gas flow). With an ambient temperature range of −20° F. to100° F., the exhaust gas flow would vary approximately 13%. For GTs inthe prior art, load changes alone could account for large changes inexhaust gas flow. The GE Model PG7241(FA) gas turbine, at 55% load,produces only 70% of full load exhaust flow. With ambient changes, thistotal flow change could be only 61% HRSG design flow. This off designflow results in inefficiency in the HRSG and requires design compromisesto accommodate such a wide range of operating conditions.

Due to the large temperatures in the HRSG as a result of continuoussupplemental firing, the LMTDs seen in several preferred exemplaryembodiments are greater than those in the prior art, Therefore, therequired surface areas are reduced and the overall size of the HRSGs maybe smaller. This results in a substantial cost savings in terms of bothconstruction and floor space costs.

HRSG Controls

In the prior art, HRSG controls for balancing the heat transfer werelimited. Desuperheating controls in the superheater and the reheaterwere common. Supplemental firing to control the steam production is nottypically used due to its negative impact on efficiency, and its addedcost. Bypasses around some economizer and feedwater sections weresometimes utilized in the prior art.

With several of the preferred exemplary embodiments, steam production isessentially controlled by the supplemental firing rate. More energyinput means more steam output. Multiple duct burner rows can be utilizedfor improved section heat transfer control. Multiple duct burner rowsallow fuel (heat) input at more than one position along the exhaust gasstream of the HRSG, and with limited heating and subsequent sectioncooling at several locations along the HRSG, serves to lower overallHRSG temperatures (possibly avoiding the more expensive water-wallconstruction).

As with the prior art, desuperheating controls will be used in thesuperheater, while desuperheating in the reheater should be limited toemergency control of reheat steam temperatures. Reheat steamtemperatures can be maintained by careful selection of the HRSG heatexchange areas and by adjustment of trim flow in a split superheaterarrangement. Feedwater flow through the HP economizer is controlled tothe optimum exhaust gas/feedwater flow ratio as is the LP economizerflow. With only one pressure level and six sections, the HRSG in thisexemplary embodiment is much simpler to control and adjust than the12-section, three pressure level boiler from the prior art illustratedin FIG. 6.

HRSG Comparison—Preferred Embodiment to Prior Art

The HRSG in several preferred exemplary embodiments may in manycircumstances be similar to the HRSG in the prior art in that it willhave a large number of tubes that transport the feedwater and recoverheat from the GT exhaust gases and transfer it to the water/steam in thetubes through convective heat transfer. This device will be very large.Both a preferred exemplary embodiment and the prior art HRSGs will becontained in a large housing that directs the GT exhaust gases from theGT exhaust to the HRSG exhaust stack. The HRSG may be oriented in eithera horizontal or vertical orientation as required to meet mechanicalconstruction constraints.

Several preferred embodiments of the present invention, however, willhave only one pressure level. This does not exclude the use ofadditional pressure levels, only that single pressure level is exemplaryof a preferred best mode exemplary embodiment. This arrangementcontrasts with the prior art which utilized multi-pressure level HRSGsto maximize heat recovery.

With only one pressure level and design for continuous supplementalfiring, a preferred exemplary HRSG embodiment may require less heatexchange area than the prior art. This will serve to reduce overallsize, footprint, weight and cost. Some of the cost savings, however,will be offset by the need for higher temperature materials and/orperhaps water-wall construction in a preferred exemplary HRSGembodiment.

With less surface area in a preferred exemplary HRSG embodiment, it islikely that the exhaust backpressure experienced by the GT due to theHRSG will be reduced. This will serve to increase the GT output andefficiency. Supplemental firing, however, tends to increase thisbackpressure and will reduce some of the performance gains achieved as aresult of lower exhaust gas restriction.

Due to the flexibility added by a preferred exemplary embodiment to thesteam cycle, the GTs will operate at full load over a wider range oftotal combined cycle plant output. This factor serves to provide a moreconstant flow to the HRSG, provide for a more optimized design, andeliminate inefficient operation at part load conditions.

With only one pressure level, the HRSG from a preferred exemplaryembodiment will be easier to monitor and control. With only smallchanges in flow and/or temperatures in the HRSG, a preferred exemplaryembodiment is able to make small adjustments in the sectionfeedwater/steam flows to compensate for these changes. With the addedsections, greater variations in exhaust gas flow, and its lesscomprehensive control system, the HRSG in the prior art was more of areactive system to the ever changing system parameters, versus apreferred exemplary embodiment which is more of a proactive system.

New Overall Combined Cycle Power Plant

The new overall combined cycle power plant of a preferred exemplaryembodiment will be similar to the prior art, but will have both subtleand major differences. The major pieces of equipment, their operation,and cost impact will now be examined and compared relative to the priorart.

Gas Turbines

The GTs utilized in several preferred exemplary embodiments may bestandard GTs as would be used in the prior art. The only differencewould be from a performance standpoint regarding the amount of pressuredrop through a preferred exemplary HRSG embodiment. The basic engine,controls, packaging, and overall arrangement may be unchanged from theprior art. Therefore, there are no engineering or development costsassociated with this major piece of equipment. This allows the use ofproven technology and helps maintain a high level of power plantreliability. Obviously, GT performance enhancements such as inletchilling, evaporative cooling, and other such methods to increase GToutput may be utilized.

HRSGs

The HRSGs from a preferred exemplary embodiment may be smaller, morecompact, single pressure level, have controlled heat transfer, and beoptimized for continuous supplemental firing. With one pressure levelversus multi-pressure levels, some preferred exemplary HRSG embodimentsmay be simpler to operate and monitor. Controls may be employed whichcontrol the firing rate, sectional flows, and/or section outlettemperature to provide optimum heat recovery and cycle efficiency for agiven set of operating parameters.

With the operational flexibility designed into the steam cycle, the GTswill be able to operate at full load over a wide range of power plantload, providing a more consistent exhaust gas flow to the HRSGs and thusmuch more efficient performance. Fewer pressure levels, higher cycleefficiency, more consistent operation, all lead to better reliabilityand lowered O&M costs.

The need for higher temperature materials or perhaps water-wallconstruction in some preferred exemplary HRSG embodiments will tend toraise the initial cost and also increase the O&M costs. It is doubtfulthat these increased costs will be more than the savings realized fromeliminating other pressure levels, associated controls, and extra heatexchange area.

HRSGs such as those illustrative of the present invention teachingscurrently do not exist in the form as described. However, conventionalsteam power plant boilers have been built for decades, and thistechnology could certainly be applied to some preferred exemplary HRSGembodiments. In addition, numerous HRSGs have been built withmulti-pressure and single pressure levels, and many have been built withsome degree of supplemental firing (including the higher temperaturewater wall construction). Of all the major components in a preferredexemplary embodiment, this one will require the most engineering anddesign effort. However, as stated previously, the continuously firedHRSG with a single pressure level is a novel concept for thisapplication, but is not beyond technological practice or capability forthose skilled in the art.

Steam Turbines

In the prior art, the STs were designed basically by the heat recoveredby the HRSG. On large combined cycle plants, a rule of thumb is that theST output is approximately 50% of the combined GT output. Withsupplemental firing this percentage could be increased, but due to thenegative effect on efficiency that was experienced utilizing thehardware, systems, and methods of the prior art, these increases weretypically small. GE informative document GER-3574F (1996), entitled “GECombined-Cycle Product Line and Performance” by David L. Chase, Leroy O.Tomlinson, Thomas L. Davidson, Raub W. Smith, and Chris E. Maslak, inTable 14 indicates that HRSG supplemental firing can increase combinedcycle plant output in the prior art by 28%, but only with an increase inoverall combined cycle heat rate (specific fuel consumption) of 9%.

The prior art focused on multi-pressure level HRSGs and STs that usedthis steam. Consequently, the STs had relatively small HP flows,moderate IP flows due to the addition of IP steam from the HRSG, andrelatively large LP flows due to the further addition of LP steam fromthe HRSG. This yields lower volumetric efficiency for the HP and IPsections of the ST, and potential exhaust end loading problems for theLP section. In addition, the steam cycles themselves were somewhatinefficient, as the IP and LP steam produced by the HRSG had lesspotential to produce work than the HP steam. Finally, the IP and LPsteam flows detrimentally add to both the ST exhaust end loading andalso the heat rejection requirements.

Due to the volumetric efficiency problems and cost/benefit ratios, theinlet pressure ratings for combined cycle plant STs has been limited toapproximately 1800 psia. As multi-pressure HRSGs have been employed,there has been no need for the use of conventional feedwater heating asthere has always been ample heat in the HRSG to provide this function.Thus, the increases in steam cycle efficiency from this efficiencyenhancement feature are not commonly applied.

The ST utilized in a preferred exemplary embodiment may be larger andmore efficient than that from the prior art. The ST utilized in apreferred exemplary embodiment can have a rating of approximately 0.75to 2.25 times (or more) than that of the total GT output. For anequivalent number of GTs and HRSGs capable of firing to 2400° F.,overall combined cycle plant capacity may be increased by a factor of2.00 or more over the prior art. This equates to a ST in several of thepreferred embodiments that can be rated at up to 4.50 times the ratingof the ST from the prior art (a ST that was associated with the sameGTs).

The ST may be similar to that from the prior art, but will likely havean increased inlet pressure rating. In addition, the ST in a preferredexemplary embodiment may utilize extraction steam fed feedwater heating,which will increase the steam cycle efficiency. With no IP or LP steamfrom the HRSG, the steam flow to the HP section of the steam turbine atrated conditions will be the maximum flow through any section. Thisincreases HP section volumetric efficiency. From this point, steam willbe extracted from the ST to various feedwater heaters, fuel preheaters,a smaller steam turbine driven BFP, and/or other plant services. Thisoperation reduces the exhaust end flow, reducing the possibility of highexhaust end loading in the ST. All these features are typical of a STthat would be used in a conventional steam power plant.

Due to its large increase in rating, (from approximately 50% of GT totalcapacity to approximately a range of 100% to over 200% of GT totalcapacity), the ST may require larger last stage blades and/or more LPsections. This represents a relatively low cost addition for capacitycompared to extra GTs, HRSGs, switchgear, transformers, foundations,etc. that would be required in the prior art to provide this extracapacity.

Other than its larger flow passing capability, higher rating, improvedefficiency, and larger blading and/or extra LP section(s), the ST mayappear similar to a ST in the prior art. It is designed typically toextract steam flow from the turbine for conventional feedwater heating,rather than admit flow to the turbine from the IP and LP HRSG sections.However, the ST would be extremely similar to a ST of similar rating andinlet conditions found in a modern conventional steam power plant.Therefore, this new combined cycle method allows for the use of moreconventional, higher efficiency ST hardware and more efficient steamturbine cycles. This maximizes the bottoming cycle efficiency, vastlyincreases capacity, and reduces overall combined cycle power plant sizeand installed cost, all without a sacrifice in reliability.

Operation

With the large amounts of supplemental firing (high ST/GT ratio), andthe ability to vary this rate of firing, several preferred exemplaryembodiments become an arrangement where the bottoming cycle is much moreindependent than in the prior art. Due to this phenomenon, and the factthat from an emissions and efficiency standpoint it is best to operatethe GTs at full load, most of the overall combined cycle plant loadvariations in a preferred exemplary embodiment are accomplished byvarying the rate of supplemental firing and subsequently the ST load,while the GTs continue to operate at or near full load. This contrastsfrom the prior art where supplemental firing was utilized to obtain onlyminor increases in plant output during peak operation, and overall plantload control was achieved mainly through load changes on the GTs.

In several preferred exemplary embodiments, at overall plant full load,the GTs will be at full load, and either the HRSG will have reached itsfiring temperature limit, or the ST will have reached its inlet pressurelimit. From this point, as plant load is reduced, supplemental firing isreduced, steam production is reduced, and subsequently the ST load isalso reduced. This process of load reduction continues until adequateflows can no longer be maintained in the HRSG.

Once adequate flows can no longer be maintained in the HRSG, the STand/or HRSG will reach an operational limit. At this point it will benecessary to decrease load on a GT or GTs. As the total GT load isreduced, ST load can be increased to meet system load. Refer to FIG. 43for a suggested mode of operation with multiple GTs. This control methodmay be used to reduce load from overall plant full load down to the HRSGand/or ST low limit by varying the rate of supplemental firing only, andallowing the GTs to operate at full load. Once at this low limit, one GTcan be unloaded, and its HRSG will begin to produce less steam.Concurrently, the remaining GT can remain at or near fall load, and itsHRSG can increase its rate of supplemental firing. This results in moresteam to the ST. The net result is a transitional zone of operationwhere one GT is reduced in load while the ST compensates for most ofthis load reduction. After reducing overall plant load sufficiently topass through this transitional zone of operation, one GT will be takenout of service (shut down), and the remaining HRSGs will be supplementalfiring at high rates and the ST will be operating at a much higher loadthan at the upper end of the transitional zone. This scheme of operationallows the GTs to remain at or near full load through a large range ofthe overall plant's expected output (approximately 50 to 100% of plantrating) with only a narrow band of operation in the transitional zonewhere one GT is brought from full load to an out-of-service condition.For FIG. 43, this transitional zone of operation is between 70% and 80%of plant load.

An exemplary embodiment of a control structure implementing the aboveprocedures is illustrated conceptually in the flowcharts of FIGS. 16-19.Discussion of this embodiment is detailed later in this document.

Performance

Since the rate of supplemental firing is large compared to the priorart, the ST capability is greatly increased. By utilizing an HRSGcapable of 2400° F. inlet temperatures, the ST can be designed (forexample) at its rated point to be approximately 2.25 times the output ofall the GTs combined. This is substantially more than a ST from theprior art, as in these applications, ST rated output was typically inthe range of 0.4 to 0.6 times the output of all the GTs. This greatlyincreases the capacity of the power plant, as the ST is now capable ofratings that are up to 4.50 times that of the ST in the prior art. Also,as previously mentioned, the operational flexibility afforded by thisarrangement allows for operation of the GTs at full load over a widerange of overall plant output. This increases the plant's part loadefficiency and lowers NOX emission levels for GTs which typicallydemonstrate increased NOX emissions at part load operation.

With this large increase in capacity over the prior art, the addedflexibility, and lowered cost per kW of capacity, this example of apreferred exemplary embodiment combined cycle plant is more adept bothoperationally and economically to provide the temporary powerrequirements of seasonal peak loads. In addition, small operationalvariables (like the isolation of feedwater heaters or operation with theHP inlet pressure at 5% over rated) will allow this example of apreferred exemplary embodiment to attain even greater capacity thanrated, but at a slight cost in efficiency. Since seasonal peaks may havedurations that last for only a matter of days each year, this is aninexpensive method to generate more power during peak periods (which maybe sold at very high rates) for minimal cost. The increased revenue isenvisioned to more than compensate for the inefficiencies and theincreased fuel costs incurred during these temporary peak loadingconditions, thus making this an economically advantageous alternativefor plant designers and electric utilities. As reported in POWERMAGAZINE, (ISSN 0032-5929, March/April 1999, page 14):

“Reserve margins are down nationwide . . . Last summer's Midwest pricespikes, up to [US]$7000/MWh [(US$7.00/kWh)], garnered most of the presscoverage, but spikes of [US]$6000/MWh [(US$6.00/kWh)], also occurred inAlberta . . .

Although it has been stated in the prior art that supplemental firingdecreases overall combined cycle thermal efficiency, this example of apreferred exemplary embodiment has shown this assumption to beincorrect. By utilizing the fuel added in supplemental firing to notonly add heat, but upgrade the bottoming cycle efficiency, it ispossible to meet or exceed prior art overall combined cycleefficiencies. This is accomplished through the use of higher inlet steampressures, larger more efficient STs, the conversion of lower pressuresteam utilized in the prior art to high-pressure steam, and the use ofconventional feedwater heating. Part load operation is also improved asthe GTs in this example of a preferred exemplary embodiment will operateat full load (where they are most efficient) for a vast majority oftheir operation (neglecting the time when they may be out of service).

Part Load Performance

As system load is reduced, the combined cycle plant load must be reducedto meet the electrical system demand. In the prior art, this wasaccomplished by a reduction in load on the GTs. This mode of loadcontrol causes a decay in the GT efficiency as well as the overallcombined cycle plant efficiency. With several of the preferredembodiments of the present invention, however, load control isaccomplished more through the variation of the amount of supplementalfiring. In this manner, the GTs remain at or near full power where theyare the most efficient and have the lowest emissions. The bulk of theload modulation is then accomplished by a reduction in the amount ofsteam production and a subsequent reduction in the output of the ST.This mode of operation provides for improved part load efficiency forthe overall combined cycle plant, as well as a reduction in maintenanceon the GTs as a result of the reduction in thermal cycling operation (GTinternal temperatures typically vary with changes in GT loading).

FIG. 33 indicates some part load efficiencies that can be expected fromconventional combined cycle power plants in the prior art and also thosethat can be attained with several of the preferred embodiments of thepresent invention. As can be seen from these curves, the prior artcombined cycles continually degrade from their optimum performance asload is reduced from 100%. However, several of the preferred embodimentsof the present invention actually experience an increase in efficiencyas load is initially reduced from 100% before it begins to degrade belowabout 80%. This part load efficiency profile for several of thepreferred embodiments of the present invention provide for substantialfuel savings as compared to the conventional combined cycle in the priorart.

Peak Load Performance

The present invention is particularly well suited for providing power atperiods of peak load. During these periods, the output of a preferredembodiment combined cycle power plant may be temporarily extended beyondits nominal rated load. As mentioned previously, this temporaryextension beyond rated power plant load may provide an enormous economicbenefit, as peak power can sell for hundreds of times the normal priceof non-peak generated electric power. Therefore, there is a strongincentive for power plant owners to generate this power. As previouslymentioned, the prior art has addressed this problem by utilizingsupplemental firing in the HRSG. Not only does this reduce theconventional combined cycle efficiency in the prior art at peak loads,but also due to the need for added ST capacity the base prior artcombined cycle efficiency is also reduced at non-peak loads as well (STis already at part load with no supplemental firing). Thus, the abilityto extend the peak power rating of conventional combined cycle powerplants comes with a detriment to the overall plant efficiency at allplant load operating points.

Since ST capacity can be increased through greater mass flow, techniquesthat increase steam flow through the ST will normally increase overallST output. Since the present invention teaches a predominantly Rankinecycle combined cycle, and as such, increases in the ST output affect awider variation in the overall combined cycle power plant capacity.Therefore, this effect to ST output is much more effective than in theprimarily GT-based combined cycle power plants as taught by the priorart.

Note in the following table that as the pressure is increased acorresponding increase in steam flow takes place. If this pressureincrease is coupled with a corresponding decrease in inlet steamtemperature, further increases in mass flow are attainable. Inconjunction with this, isolation of feedwater heaters will serve todirect more steam flow to the exhaust of the ST, further increasing SToutput. Unlike the prior art, this method provides the ability to extendthe peak power rating of combined cycle power plants implementing thepresent invention without incurring a detriment to the overall plantefficiency at non-peak plant load operating points.

Peak Power Extension Example Inlet Tem- Mass/ Inlet perature SpecificSteam Volume Steam Flow Pressure (degrees Volume Flow Flow Increase(psia) F) (ft³/lb) (lb/hr) (ft³/hr) (new/old) 2400 1050 0.338245 2000000676490 Baseline 2520 1050 0.320349 2111730 676490 1.055865 2520 10000.304021 2225143 676490 1.112571 2520 950 0.286872 2358162 6764901.179081 2640 1000 0.288236 2346998 676490 1.173499 2640 950 0.2715542491183 676490 1.245591

Cost

A substantial advantage to this exemplary preferred embodiment is thecost savings. As mentioned previously, a plant with HRSGs designed forup to 2400° F. inlet temperature through supplemental firing can easilyhave a ST rated 2.0 times the total GT capacity. Therefore, total plantoutput is 3.0 (2.0ST+1.0 GT) times the GT capacity. A combined cycleplant from the prior art would have a ST rated at approximately 0.5times the total GT capacity. Therefore, the capacity ratio isessentially (3.0/1.5)=2.0. In other words, the combined cycle plant fromthis preferred exemplary embodiment will have 100% greater capacity thanthe prior art. An example of this trend is demonstrated in FIG. 39,which is a heat balance for a 1040 MW exemplary preferred embodimentutilizing two (2) industry standard GE model PG7241FA GTs and a largeST. FIG. 22 illustrates a combined cycle from the prior art utilizingthe same quantity and model of GTs and the standard smaller ST,nominally rated at 520 MW.

This means that to provide capacity equal to that from this example, acombined cycle plant from the prior art would need to add 100% moreequipment. This means more GTs, another ST, more HRSGs, switchgears,transformers, and all the necessary systems and real estate required tosupport this equipment. This will serve to raise the plant installedcost by essentially 100%.

In terms of 1999 dollars, a modern high efficiency large combined cyclepower plant could be installed for approximately US$450 per kW ofcapacity. Therefore, a 720 MW plant (720,000 kW) would cost US$324million to construct. If this plant were to be expanded to 1050 MW, theinstalled cost would climb to US$472 million. In contrast, the presentinvention teaches that it is possible to use less equipment to affectthis expansion, thus decreasing the cost per kW of rated plant capacity.

Retrofits

Another prime application for this example of a preferred exemplaryembodiment is in retrofit applications of existing plants. Manysteam-powered plants in existence today will produce expensive powercompared to the highly efficient combined cycle plants discussed herein.With electrical deregulation on the horizon, it will be imperative thatpower producers be competitive. Therefore, technology that will helpexisting steam plants compete with new combined cycle plants is needed.

Since this example of a preferred exemplary embodiment operates(predominantly) on a single pressure level, utilizes higher steampressures that are typical for STs found in conventional steam plants,has a higher ST/GT output ratio, and provides for a compact design, itis ideally suited for retrofit applications of existing steam powerplants. With a preferred exemplary embodiment, large steam plants couldactually bypass their existing boilers and utilize steam directly fromthe HRSGs. This increases cycle efficiency and (in many cases) wouldreduce plant emissions drastically. This could be accomplished using theexisting ST, condenser, and other infrastructure already in the existingplant. This would provide the owners with a highly efficient combinedcycle plant with reduced capital investment and minimal real estaterequirements.

Exemplary Preferred Embodiments—Typical Configuration

Overview

The configuration of several of the preferred embodiments is similar tothe prior art, in that GTs and HRSGs are utilized to produce power andconvert exhaust gas heat into steam. However, several of the preferredembodiments will utilize a continuously fired HRSG that producessignificantly more steam, and do so at a single pressure level (orprimarily a single pressure level). This higher quantity (and typicallyhigher pressure) steam drives a ST that is much larger in comparison tothe ST in the prior art that was associated with the same GTs.

Due to the large feedwater flows, feedwater will be heated in the HRSGsas well as in a separate feedwater heating loop which utilizesconventional ST extraction steam fed feedwater heaters. Fuel gas heaterswill also be employed to improve cycle efficiency.

Embodiment of FIG. 9

Design

Refer to FIG. 9 for a schematic representation of one exemplarypreferred power plant embodiment utilizing the teachings of the presentinvention. The GTs (920) each exhaust into their respective HRSGs anddrive their respective generator (921). These exhaust gases producesteam in the HRSG that subsequently produces power in the ST and whichis ultimately condensed in the condenser (939).

Feedwater Heating—HRSG Feedwater Heating Loop

Condensate from the condenser (939) goes to the LP-BFP (930) where it ispumped to an intermediate discharge pressure. From here, the LPfeedwater control valve (960) maintains an optimum flow through the LPeconomizer (901) while diverting the excess feedwater flow to theconventional feedwater heater(s) (933). Flow exiting the LP economizercontinues to the HP-BFP (931) and is pressurized to a pressure that isequal to inlet steam pressure plus an allowance for pressure drops inthe system. From here it flows through the HP economizer (902) and(903). Some feedwater flow, however, after exiting the LP economizer, isdiverted through the feedwater balancing valve (967) so as to maintainan optimum flow through the HP economizer sections (902) and (903). Thediverted feedwater that passes through the feedwater balancing valve(967) combines with the feedwater exiting feedwater heater (933). Thiscombined flow now continues to the second HP-BFP (932) where it ispressurized to a pressure similar to that of HP-BFP (931). The divertedfeedwater flow exiting HP-BFP (932) goes to feedwater heater(s) (934).The feedwater exits feedwater heater(s) (934) and combines with thefeedwater flow exiting the HP economizer. This flow is now available atdesuperheating lines (950) and (951), while the bulk of the flowcontinues to the evaporator section (904).

Evaporator

In the evaporator section, the feedwater is boiled into steam andtravels to the superheater section (905). If the superheated steam istoo hot, condensate is sprayed through line (950) into the superheatersupply line to control the HP turbine section (935) inlet temperature tothe desired temperature. Steam expands in the HP turbine section down tothe exhaust point, and becomes known as cold reheat steam. The coldreheat steam continues to the reheater section (906) in the HRSG.

Reheater

On its way to the reheater section, some steam passes through non-returnvalve (964) to line (954). This steam travels to the feedwater heater(934), which preheats the feedwater flowing through same. The condensedsteam from this feedwater heater cascades down to the inlet of theHP-BFP (932).

The cold reheat steam from the HP turbine section exhaust now travelsthrough the reheater section of the HRSG (906) for return to the IPsection of the ST. If its temperature is too high, condensate is sprayedthrough line (951) into the reheater supply line to control the IPturbine section (936) inlet temperature to the desired temperature.Steam expands in the IP turbine section down to its exhaust point, andbecomes known at this point as crossover steam.

Crossover Steam

The crossover steam continues to the LP sections (937) of the ST. On itsway to the LP section, some crossover steam is diverted throughnon-return valve (965) to line (955). This steam travels to feedwaterheater (933), which preheats the feedwater flowing through same. Thecondensed steam from this feedwater heater flows to the outlet of thecondenser.

Steam expands in the LP turbine sections and exhaust into the condenser(939). Shaft horsepower produced in the ST drives the generator (938),which produces electrical power.

Note that in this example cold reheat and crossover steam is used toprovide extraction steam to the feedwater heaters. Although these aretraditional points for the supply of this steam, this does not precludethe utilization of extraction steam from any practival point on the STto provide this function.

Low Load Operation

For operation at low loads, there is insufficient HP steam flow tomaintain optimum levels of feedwater through the HRSG. In this mode ofoperation, valves (960) and (967) are closed. With no feedwater flow toremove heat, all extraction lines (954, 955) pass no flow. All feedwaterflow, therefore, passes through the HRSG as the parallel feedwater loopis closed off.

As load is decreased from this point by a reduction in steam flow(reduction in supplemental firing), the feedwater flow through the HRSGis no longer sufficient to absorb the exhaust gas heat and yet stillmaintain optimum exhaust gas temperature. Therefore, operation belowthis point will result in increased exhaust gas temperatures and lowercombined cycle efficiency. At this point, the design engineer will needto evaluate performance parameters and determine if it is moreeconomical at this point of operation to reduce load of the GTs, orcontinue modulating supplemental firing rates and allowing the HRSGexhaust gas temperature to increase. At some point of reduced load,however, it will become economically favorable to reduced load on theGTs.

Embodiment of FIG. 15

Design

Refer to FIG. 15 for a schematic representation of another exemplarypreferred power plant embodiment utilizing the teachings of the presentinvention. The GTs (1520) each exhaust into their respective HRSGs(1509) and drive their respective generator (1521). These exhaust gasesproduce steam in the HRSG that subsequently produces power in the ST andwhich is ultimately condensed in the condenser (1595).

Feedwater Heating—HRSG Feedwater Heating Loop

Condensate from the condenser (1595) goes to the LP-BFP (1530) where itis pumped to its discharge pressure. From there, the LP feedwatercontrol valve (1560) maintains an optimum flow through the LP economizer(1501) while diverting the excess feedwater flow to the first of aseries of conventional feedwater heaters (1533). Flow exiting the LPeconomizer continues to the HP-BFP (1531) and is pressurized. From hereit flow through the HP economizer (1502). However, after exiting the LPeconomizer some feedwater flow is diverted through the feedwaterbalancing valve (1561) so as to maintain an optimum flow through the HPeconomizer section (1502). In addition, some flow is diverted to thefuel gas heater (1575) through line (1571). After pre-heating the fuelgas, this flow is returned to the inlet of the LP-BFP (1530) via line(1572). The remaining feedwater continues to the HP economizer, and flowexiting the HP economizer combines with the feedwater flow exiting thefinal feedwater heater (1537). This flow is now available atdesuperheating valves (1510) and (1511), while the bulk of the flowcontinues to the evaporator section (1504).

Feedwater Heating—Conventional Feedwater Heating Loop

In the parallel feedwater heating loop, feedwater proceeds through thefirst feedwater heater (1533) where it is heated. This flow then travelsthrough the second and third feedwater heaters (1534) and (1535)respectively. At the exit of feedwater heater (1535), flow diverted fromthe HRSG parallel loop through the feedwater balancing valve (1561)combines with this feedwater and continues to a HP-BFP (1532) where itis pressurized. From here it travels through the fourth and fifthfeedwater heaters (1536) and (1537) respectively. The feedwater fromthis heating loop now combines with the feedwater from the HRSG parallelloop and is fed to the evaporator section (1504) of the HRSG (minus flowrequired by the desuperheating valves (1510) and (1511)).

Evaporator

In the evaporator section, the feedwater is boiled into steam andtravels to the superheater section (1505). If the superheated steam istoo hot, desuperheating valve (1510) modulates to spray condensate fromthe desuperheating line (1550) into the superheater supply line andcontrol the HP turbine section (1590) inlet temperature. Steam expandsin the HP turbine section until reaching the first extraction where asmall portion of the steam is removed from the turbine throughnon-return valve (1568) to line (1558). This steam is fed to the fifthfeedwater heater (1537) which preheats the feedwater flowing throughsame. The condensed steam from the fifth feedwater heater cascades downto the fourth feedwater heater (1536). The steam in the HP section ofthe ST (1590) that is not extracted continues to the section exit point,and becomes known as cold reheat steam. The cold reheat steam continuesto the reheater section (1506) in the HRSG.

Reheater

On its way to the reheater section, some steam (second extraction)passes through non-return valve (1564) to line (1554). This steamtravels to the fourth feedwater heater (1536) which preheats thefeedwater flowing through same. The condensed steam from the fourthfeedwater heater cascades down to the inlet of the HP-BFP (1532).

The cold reheat steam now travels through the reheater section of theHRSG for return to the IP section of the ST. If its temperature is toohigh, desuperheating valve (1511) modulates to spray condensate from thedesuperheating line (1551) into the reheater supply line and control theIP turbine section (1591) inlet temperature. Steam expands in the IPturbine section until reaching the third extraction where a smallportion of the steam is removed from the turbine through non-returnvalve (1567) to line (1557). This steam is fed to the third feedwaterheater (1535) which preheats the feedwater flowing through same. Thecondensed steam from the third feedwater cascades down to the 2^(nd)feedwater heater. The steam in the IP section of the ST (1591) that isnot extracted continues to the section exit point, and becomes known ascrossover steam.

Crossover Steam

The crossover steam continues to the LP sections (1592) and (1593) ofthe ST. On its way to the LP section, some steam (fourth extraction) isdiverted through non-return valve (1565) to line (1555). This steamtravels to the second feedwater heater (1534) which preheats thefeedwater flowing through same. The condensed steam from the secondfeedwater heater cascades down to the first feedwater heater (1533).

Steam expands in the LP turbine sections until reaching the fifthextraction where a small portion of the steam is removed from theturbine through non-return valve (1569) to line (1559). This steam isfed to the first feedwater heater (1533) which preheats the feedwaterflowing through same. The condensed steam from the first feedwaterheater is returned via line (1512) to the inlet of the LP-BFP (1530).

The steam in the LP sections of the ST (1592, 1593) that is notextracted continues through the section to exit at the condenser (1595).Shaft horsepower produced in the ST drives the generator (1594) whichproduces electrical power.

Low Load Operation

For operation at low loads, there is insufficient HP steam flow (thuslow flows of condensate from condenser) to maintain optimum levels offeedwater through the HRSG. In this mode of operation, valves (1560) and(1561) are closed. With no feedwater flow to remove heat, all extractionlines (1558, 1554, 1557, 1555, 1559) pass no flow. All feedwater flow,therefore, passes through the HRSG as the parallel feedwater loop isclosed off.

As load is decreased from this point by a reduction in steam flow(reduction in supplemental firing), the feedwater flow through the HRSGis no longer sufficient to absorb the exhaust gas heat and yet stillmaintain optimum exhaust gas temperature. Therefore, operation belowthis point will result in increased exhaust gas temperatures and lowercombined cycle efficiency. At this point, the design engineer will needto evaluate performance parameters and determine if it is moreeconomical at this point of operation to reduce load on the GTs, orcontinue modulating supplemental firing rates and allowing the HRSGexhaust gas temperature to increase. At some point of reduced load,however, it will become economically favorable to reduce load on theGTs.

Exemplary Preferred Embodiment—725 MW Power Plant

Overview

As an example of another preferred exemplary embodiment, a 725 MWnominal capacity combined cycle power plant design will be examined.This exemplary power plant will utilize two (2) GE Model PG7241(FA) GTs.These GTs will each exhaust into its own single pressure HRSG designedfor 2400 psia operation. A nominal 400 MW reheat ST will be usedexhausting to a once through condenser operating at 1.2 inches HgA(inches of mercury absolute) exhaust pressure. Due to the largefeedwater flows, feedwater will be heated in the HRSGs as well as in aseparate feedwater heating loop which utilizes conventional STextraction steam fed feedwater heaters. Fuel gas heaters will also beemployed to improve cycle efficiency.

Design

The GE GT design is rated 170,770 kW based upon ISO conditions, with a3.0 inches of H₂O inlet air pressure drop and 10.0 inches H₂O exhaustgas pressure drop through the HRSG. Total GT output is therefore 341,540kW. Refer to FIG. 35 for a schematic representation of this exemplarypower plant. The numbers indicated at various points along the processcorrespond to “point” numbers tabulated in FIGS. 36, 37, and 38. Thedata corresponding to the “point” numbers tabulated in FIG. 36, FIG. 37,and FIG. 38 identifies the pressure, temperature, enthalpy, and flow atthe corresponding “point” . This overall information contained in FIGS.35-38 represents what is termed a “heat balance”, which is an overallenergy and mass balance for the cycle. Note for this example, deaerationis completed in the condenser.

Layout

FIG. 26 illustrates the physical plant layout of this example of severalof the preferred embodiments. Note that it is extremely similar to theGE S207FA combined cycle power plant in the prior art, shown in FIG. 22.The most noticeable difference between the two layouts is theconfiguration of the ST. In the prior art, the ST has a relativelyunderutilized HP/IP section, and one LP section. In several of thepreferred embodiments, the HP/IP section is similar to the prior art,but has considerably increased volumetric flow. To efficiently use thehigher steam flows at lower pressure, a second LP section is shown.However, this second section may not be required, depending upon theeconomic evaluation.

Comparison to Prior Art

FIG. 22 and FIG. 24 are layouts of the GE S207FA combined cycle and theWestinghouse 2X1 501G combined cycle power plants respectively. The GEfacility requires approximately 2.3 acres of real estate while theWestinghouse facility requires approximately 3.3 acres. The powerdensity is nearly the same for these two options at 220 MW per acre.Several of the preferred embodiments, however, can be designed as shownin FIG. 26 to be 725 MW as in the example, which is 315 MW per acre, orit can be designed for up to 1050 MW (see FIG. 29) which is 455 MW peracre. This allows for the production of significantly more power withonly a given amount of real estate. This factor is advantageous for newconstruction, but will also be especially appreciated for retrofit ofexisting plants where real estate comes at a premium.

Besides the premium for real estate, the combined cycles in the priorart are also more expensive from a fuel consumption, capital cost, andmaintenance perspective. FIG. 23 and FIG. 25 are economic pro forma forthe GE S207FA combined cycle and the Westinghouse 2X1 501G combinedcycle power plants respectively. These figures tabulate the annual costsfor fuel, capital, and maintenance for each power plant. FIG. 27 is theeconomic pro forma for an exemplary preferred embodiment of the presentinvention. Note that each individual cost for fuel, capital, andmaintenance is less than the each individual cost for combined cyclepower plants from the prior art. Therefore, the cost to produceelectricity is reduced in all major cost categories by several of thepreferred embodiments.

Exemplary Preferred Embodiment—Supercritical Steam Conditions

Overview

As another example of a preferred exemplary embodiment, a 1040 MWnominal capacity combined cycle power plant design utilizingultrasupercritical steam conditions with elevated steam temperatureswill be examined. This exemplary power plant will utilize two (2) GEModel PG7241(FA) GTs. These GTs will each exhaust into its own singlepressure HRSGs designed for 3860 psia operation. A nominal 730 MW doublereheat ST will be used exhausting to a once through condenser operatingat 1.2 inches H_(g)A (inches of mercury absolute) exhaust pressure. Dueto the large feedwater flows, feedwater will be heated in the HRSGs aswell as in a separate feedwater heating loop which utilizes conventionalST extraction steam fed feedwater heaters. Fuel gas heaters will also beemployed to improve cycle efficiency.

Design

The GE GT design is rated 168,815 kW based upon ISO conditions, with a 3inches of H20 inlet air pressure drop and a 10.0 inches H2O exhaust gaspressure drop through the HRSG. Total GT output is therefore 341,540 kW.Refer to FIG. 39 for a schematic representation of this exemplary powerplant. The numbers indicated at various points along the processcorrespond to “point” numbers tabulated in FIGS. 40, 41, and 42. Theremaining data corresponding to the “point” numbers tabulated in FIG.40, FIG. 41, and FIG. 42 identifies the pressure, temperature, enthalpy,and flow at the corresponding “point”. This “heat balance” is an overallenergy and mass balance for the cycle. Note for this example, deaerationis completed in the condenser.

Comparison to Prior Art

The elevated steam temperatures (1112° F.) and pressures (3860 psig) areindicative of those used in advanced steam cycles, sometimes referred toas ultrasupercritical. Refer to the informative document entitled “SteamTurbines for Ultrasupercritical Power Plants” by Klaus M. Retzlaff andW. Anthony Ruegger (General Electric Reference GER-3945, 1996) forinformation on ultrasupercritical steam turbines and their cycles. Notethat at an exhaust temperature of 1123° F., the industry standardGeneral Electric (GE) Model PG7241(FA) Gas Turbine does not havesufficient high temperature exhaust energy to produce these steamtemperatures at the required flows. Therefore, such conditions were noteven available in the prior art.

The supercritical steam power plant of the preferred embodiment of thepresent invention is similar to the subcritical steam power plant of thepreferred embodiment, with the primary difference being improvedefficiency. Greater steam pressures, higher steam temperatures, and theuse of the second reheat provides the added efficiency for thisapplication. Note that with these steam conditions, and even with thelarge extension in capacity (100%) the combined cycle efficiency for thepreferred embodiment of the present invention approaches that of theprior art with the same technology GTs (6229 BTU/kWh versus 6040BTU/kWh).

However, efficiency is only one part of the economic equation. The othermajor costs, capital expenditure and maintenance, will be greater withthe supercritical preferred embodiment versus subcritical. Therefore, aspreviously discussed, a total economic analysis must be completed todetermine the optimum arrangement for an individual preferred embodimentcombined cycle power plant. In general, when fuel costs are high,supercritical applications will become the economic optimum, and whenfuel costs are low, subcritical applications will be preferred.

Power Plant Load Profile

Dispatched Power Plants

As previously discussed, to maintain a constant frequency of power (60Hz in the US), the power produced by all power plants connected to thegrid must equal the power being consumed by the users on the grid.Therefore, power plants have their output “dispatched”, or controlled bythe Power Pool to meet the system demand.

As a result of being dispatched, most power plants will spend verylittle of their operational time at rated output. Instead of operatingat full rated capacity, most power plants will operate at someintermediate load and share the system load with all other power plantsconnected to the grid. This statistic may be visually confirmed byinspecting the load duration curve of FIG. 31B, which represents atypical long-term distribution of utilized plant load versus percentageof time. Note that using this long-term data, most power plants willoperate at peak load less than 10% of the time, and will be atintermediate load levels for 70% of the time.

FIG. 31A provides typical hourly load data for the South Atlantic Regionof the U.S. over a 24-hour period. As can be seen from this data, thepeak load of 62,000 megawatts (MW) for the day is substantially higherthan the low of 40,000 megawatts. In addition, the total system capacityis likely higher than 62,000 MW, perhaps 70,000 MW (70 gigawatts, GW).This means that except for seasonal peaks (i.e. hot summer days), evenduring non-seasonal peak hours, many power plants are not operated atrated capacity. Therefore, dispatched power plants can expect to seelarge load variations and potentially spend only a matter of hoursannually at rated capacity.

To determine a typical conservative load profile, the data from FIG. 31Awas blocked into segments. The periods when the load was above 60 GW wasdetermined to be peak operation. The periods of operation between 50 and60 GW was considered to be intermediate power operation, and periodsbelow 50 GW were considered to be night operation. This profile wasconsidered to be an average weekday. For weekends, 8 hours per day wasconsidered intermediate, while the remainder was taken to be nightoperation (using weekday averages for intermediate and night power onweekends). FIG. 32 provides the details of these calculations. Utilizingthe data calculated in FIG. 32, a typical load profile to be used forcomparison purposes is as follows:

Period Average Plant Capacity (%) Hours Per Week Night 60 77Intermediate 80 71 Peak 100 20

Note that although the capacity per FIG. 32 for peak is only 87.86%,this number has been adjusted to 100.00% for discussion purposes. Thenight and intermediate capacity numbers have been adjusted by less than1% from the values in FIG. 32, and are adjusted downward to compensatefor the upward adjustment to peak operation.

Exemplary Power Plant Load Profile

Utilizing the data from the above table, the calculated load profile canbe used for the purpose of determining an annual capacity factor andquantity of fuel consumed for a given combined cycle power plant, basedupon part load operation data in FIG. 33. It is significant to note fromthe table above and FIG. 31B that the plant efficiency using the priorart technology will rarely (if ever) reach optimum economic performance.In contrast, the present invention embodiments as illustrated in FIG. 33will always be more optimal than the prior art configurations.

Economics of the Present Invention

Economic Considerations

The costs for operating a combined cycle power plant are varied.However, the three largest costs for the power plant operators typicallyare fuel, capital cost (debt), and maintenance. These three costsconstitute the major portion of the cost (expressed in $/kWh) to produceelectricity at large combined cycle power plants. Some of the minorcosts include payroll for the operations staff, taxes, insurance,license fees, and other miscellaneous expenses. For an economiccomparison of several of the preferred embodiments of the presentinvention to the prior art, focus will be on the three major expenses:fuel, debt, and maintenance.

Fuel Costs

The largest cost that typically is incurred by a large, modern, combinedcycle facility is the cost of fuel. Whether the fuel is natural gas,fuel oil, or some other combustible fuel, the combined cycle facilitymust consume large quantities of fuel to produce large quantities ofelectricity. In essence, a power plant actually converts energy in oneform (raw fuel), into energy of another form (electricity). Therefore,since the function of a power plant is to perform this conversionprocess, the efficiency of this conversion process is the key to thepower plant's economic success.

Prior art combined cycle power plants have efficiencies in the generalrange of 48% (LHV) for an older design such as a GE S106B combined cycleup to 60% (LHV) for the proposed GE S107H advanced cycle which has notyet seen commercial service. These efficiencies are based upon the lowerheating value (LHV) of the fuel. However, these efficiencies are forfull load operation, and as noted in FIG. 31A and FIG. 31B, most powerplants actually spend little time at full load. For part load operation,FIG. 8 provides an indication of the efficiency loss that can beexpected at reduced loads for combined cycle power plants in the priorart. Utilizing this data, FIG. 33 illustrates the dramatic improvementin part load efficiency that is realized by several of the preferredembodiments of the present invention as compared to the combined cyclein the prior art (here a lower heat rate indicates more optimalperformance). This part load efficiency improvement, along with improvedefficiency at full load, enables several of the preferred embodiments ofthe present invention to be more economical than the prior art in termsof fuel consumption.

Based upon the load profile in FIG. 32, and utilizing the heat rate(efficiency) data from FIG. 33, FIG. 34 tabulates the annual fuel costsfor this exemplary combined cycle power plant of the preferredembodiment versus current state-of-the-art combined cycle power plantsin the prior art. In either case, many of the exemplary combined cyclepower plants of the preferred embodiment use less fuel on an annualbasis than either of the prior art combined cycles.

Capital Costs

Next to fuel costs, the most significant cost for a new combined cyclepower facility is capital cost. This is the amount of money required toservice the debt (loan payments). Although plant efficiency isimportant, the overall cost of the power plant is also an importanteconomic consideration. As discussed prior, just as the economics ofsmall portions of the combined cycle plant must be evaluated (i.e.larger ST exhaust sections), the economics of the overall combined cyclepower plant must also be evaluated. Minor decreases in plant heat rate(minor increase in efficiency) must not be more than offset by increasesin capital cost. Therefore, the power plant developers and engineersstrive to construct the best economic alternative that is available.

Due to its higher power density, utilization of less equipment, andreduced construction costs, several of the preferred embodiments of thepresent invention have significantly lower capital costs (up to a 30%reduction) than combined cycles in the prior art. Again, FIG. 34tabulates the capital costs for this exemplary combined cycle powerplant of the preferred embodiment versus current state-of-the-artcombined cycle power plants in the prior art. In either case, manyexemplary combined cycle power plants of the preferred embodimentrequire significantly less capital than either of the prior art combinedcycles.

Maintenance Costs

Another large expense for power plant owners is average annualmaintenance costs, especially maintenance costs for the large pieces ofequipment. For a large 725 MW plant in the prior art, as shown in theexample, these costs can exceed $10 million annually. Therefore, powerplants with reduced maintenance costs are economically advantageous.

By utilizing a high power density design which reduces the amount ofmajor equipment, and by utilizing low maintenance STs as the major powerproducing machines instead of high maintenance GTs, several of thepreferred embodiments of E the present invention have appreciably lowermaintenance costs than combined cycles in the prior art. In FIG. 34maintenance costs for this exemplary combined cycle power plant of thepreferred embodiment versus current state-of-the-art combined cyclepower plants in the prior art are tabulated. In either case, theexemplary combined cycle power plant of the preferred embodiment is lessmaintenance intensive than either of the prior art combined cycles.

Overall Cost Comparison

FIG. 34 provides an economic comparison of the exemplary combined cyclepower plant of a preferred embodiment of the present invention incontrast to state-of-the-art combined cycle power plants in the priorart. As can be seen from the data, this exemplary combined cycle powerplant of the preferred embodiment is less expensive to operate thancombined cycles in the prior art in all three of the major costcategories: fuel, capital expenditures, and maintenance.

In addition, compared to the Westinghouse 2X1 501G combined cycle powerplant, NOX emissions are reduced by a factor of more than three, or byapproximately 180 tons/yr. For a 20-year plant operational life, theexemplary combined cycle power plant of the illustrated preferredembodiment saves US$469 million as compared to the Westinghouse model501 G combined cycle from the prior art. These savings are more than theinitial plant construction costs of US$340 million for the Westinghouse2X1 501G combined cycle power plant, and represent a significanteconomic advantage for power producers in a deregulated, competitiveenvironment.

Operation of the Present Invention

Exemplary HRSG Control Method

Due to the unique arrangement of equipment, the use of a predominantlysingle pressure level HRSG, and the need to optimize heat recovery, anexemplary control system to meet these objectives is illustrated in FIG.16. The control system is exemplary of a combined cycle described in thepreferred embodiments illustrated in FIG. 9 and FIG. 15, although it mayhave a wide application to other embodiments of the present invention.There is one HRSG for each GT in this example. Note that this is anexample of an HRSG control system for this particular application, andis a demonstration of the principles in flow management, optimum heattransfer, and integration of HRSG and feedwater heating loops. For otherapplications, this arrangement could be modified for the particularcircumstances. However, many of the principles outlined in this controlschematic would be employed.

In FIG. 16, the control begins at (1601) and continues to process block(1602) where the loop control begins. Control then flows to processblock (1603). At this point the controller examines inputs from processblock (1611) which include ambient temperature and GT load (inparticular, the GT exhausting into the HRSG in this control loop). Basedupon a characteristic curve programmed into the software, the controllerdetermines the GT exhaust flow.

Utilizing the DCS inputs for ST required steam flow and steam flowalready being produced by the other HRSGs, at process block (1604) thecontroller calculates its required steam flow as the ST required flowminus flow from other HRSGs. Control proceeds to decision block (1605)and compares the HRSG required steam flow to the optimum flow for the HPeconomizer.

If the power plant is operating at reduced load, control flows toprocess block (1606). At this point of operation, there is less than theoptimum HP economizer flow required from the HRSG. Therefore, more heatwill be available in the GT exhaust gases than can be recovered in theHRSG. As a first phase of load reduction, the controls will begin tomodulate valves (960) and (967) in a closing direction to reduce flowthrough the parallel feedwater heating loop. Once the parallel feedwaterheating loop has been completely isolated, the second phase of controllowers the power output of the GT. Control now returns to the initialprocess block (1602).

From decision block (1605), if the HRSG required flow is greater thanthe HP economizer optimum flow, then control proceeds to process block(1620). If the GT is operating at less than full load, the first phaseof control is to increase GT load. Once the GT is operating at fullload, valve (967) is modulated to begin feedwater heating in theparallel loop. Utilizing inputs from the DCS at process block (1610) forthe evaporator section pressure and the temperature exiting the HPeconomizer, valve (967) is modulated to obtain the optimum watertemperature at the exit of the HP economizer. Pump (932) beginsoperation once flow begins to pass through valve (967).

Control now proceeds to decision block (1621). If the HRSG requiredsteam flow is less than the LP economizer optimum flow, then controlproceeds to process block (1622). At this power plant load, there isstill no need for LP feedwater heating as there is more than sufficientheat available in the exhaust gases to heat the feedwater in the LPeconomizer. Therefore, valve (960) is closed. Control returns to theinitial process block (1602).

From decision block (1621), if the HRSG required steam flow is greaterthan the LP economizer optimum flow, then control proceeds to processblock (1623). At this power plant load, conventional LP feedwaterheating is required as there is insufficient heat available in theexhaust gases to heat the feedwater in the LP economizer. Therefore,valve (960) is modulated to control flow through the LP economizer toits optimum. Control returns to the initial process block (1602).

Exemplary Overall Power Plant Control Method

In providing a control logic for the overall plant, some of the majorobjectives include improved efficiency and continuous low emissionlevels. These objectives are best attained by operating the GTs at ornear full load. The control logic for the overall combined cycle controlin this example will focus on these objectives. Obviously, one skilledin the art will recognize that to achieve other objectives, this controlscheme may be easily modified to support other priorities.

Main Control Loop

Referencing FIG. 17, the control starts at (1701) and continues toprocess block (1702) where the loop control begins. Control then flowsto process block (1704). At this point the controller examines inputsfrom process block (1703) which include the current overall plant loadand the load reference (desired plant load). Based on these inputs, thecontroller determines the load change requirements. At decision block(1705) the controller examines the need for a change in load. If thereis no need to change load, the control is returned to the initialprocess block (1702).

If a load change is required, control flows to decision block (1706)where it must be determined whether the overall plant load needs to beincreased or decreased. If it needs to be increased, process controlproceeds to the Increase Power Output subroutine (1708). An exemplaryembodiment of this subroutine is illustrated in the flowchart of FIG.18. If it needs to be decreased, process control proceeds to theDecrease Power Output subroutine (1707). An exemplary embodiment of thissubroutine is illustrated in the flowchart of FIG. 19.

Increase Power Output

Referencing FIG. 18, the Increase Power Output subroutine begins at step(1801) and proceeds to decision block (1802). If the plant is notoperating in a transition zone of operation (zone where one GT is in theprocess of either being brought into or out of service), then processcontrol flows to decision block (1804). Note that in FIG. 43, thetransition zone of operation is between 70% and 80% of plant load. Thiszone range may be varied by one skilled in the art to achieve a varietyof plant performance objectives.

If the plant is operating in a transition zone of operation, thenprocess flows to the Transition Control subroutine, (1805). An exemplaryembodiment of this subroutine is illustrated in the flowchart of FIG.20. Control then returns to the end subroutine process block (1803). Allprocess returns to this block (1803) are returned to subroutine block(FIG. 17, 1708), and finally to the initial process block for overallplant control (FIG. 17, 1702).

At decision block (1804), if all of the plant's GTs are operating, thenprocess flow proceeds to decision block (1820). At this juncture, thecontroller determines if all of the GTs are operating at fall load.Since the best method to achieve the objectives is to operate the GTs atfull load, if all GTs are not at full load, control flows to processblock (1821) where load is increased on one or more GTs. Control nowreturns to the end subroutine process block (1803).

From decision block (1820), if all GTs are at fall load, then controlflows to decision block (1822). This block determines whether or noteither the ST or HRSG is operating at an upper limit. For the HRSG, thisis typically the supplemental firing temperature. For the ST, this wouldtypically be the inlet pressure. This could also be an operational limitbased upon efficiency or another parameter. If any of these limits isreached, control flows to process block (1823) which will energize astatus light in the control room indicating to the operators that theplant is at full capacity. Control now returns to the end subroutineprocess block (1803).

From decision block (1822), if the ST or HRSG is not at an upper limit,then control flows to process block (1824), where the fuel flow to theHRSGs is increased. Control now returns to the end subroutine processblock (1803).

From decision block (1802), if all of the plant's GTs are not operating,then process flow proceeds to decision block (1810). At this juncture,the controller determines if all of the GTs that are currently operatingare at full load. Again, since the best method to achieve the objectivesis to operate the GTs at full load, if all GTs are not at full load,control flows to process block (1811) where load is increased on one ormore GTs. Control now returns to the end subroutine process block(1803).

From decision block (1810), if all operating GTs are at full load, thencontrol flows to decision block (1812). This block determines whether ornot either the ST or HRSG is operating at an upper limit. In addition toa temperature or pressure limit, this could also be an operational limitbased upon power plant efficiency or other system requirements. If anyof these limits are reached, control flows to the Transition Controlsubroutine, process block (1813). An exemplary embodiment of thissubroutine is illustrated in the flowchart of FIG. 20. Control thenreturns to the end subroutine process block (1803).

From decision block (1812), if the ST or HRSG is not at an upper limit,then control flows to process block (1814), where the fuel flow to theHRSGs is increased. Control now returns to the end subroutine processblock (1803).

Decrease Power Output

Referencing FIG. 19, the Decrease Power Output subroutine begins at(1901) and proceeds to decision block (1902). If the plant is notoperating in a transition zone of operation (zone where one GT is in theprocess of either being brought into or out of service), then processcontrol flows to decision block (1904). Note that in FIG. 43, thetransition zone of operation is between 70 and 80% of plant load.

If the plant is operating in a transition zone of operation, thenprocess flows to the Transition Control subroutine, (1905). An exemplaryembodiment of this subroutine is illustrated in the flowchart of FIG.20. Control then returns to the end subroutine process block (1903). Allprocess returns to this block (1903) are returned to subroutine block(FIG. 17, 1707), and finally to the initial process block for overallplant control (FIG. 17, 1702).

At decision block (1904), if all of the plant's GTs are operating, thenprocess flow proceeds to decision block (1920). At this juncture, thecontroller determines whether or not either the ST or HRSG is operatingat a lower limit. For the HRSG and ST, these limits would be determinedby the plant engineers who would specify the optimum point to beginshutdown of a GT. If neither of these limits is reached, then controlflows to process block (1921), where the fuel flow to the HRSGs isdecreased. Control now returns to the end subroutine process block(1903).

From decision block (1920), if the GT or HRSG is at a lower limit ofoperation, then process control proceeds to decision block (1922). Ifthe plant output is greater than the upper limit of the transition zoneof operation, control flows to process block (1924) where load isdecreased on one or more GTs. Control now returns to the end subroutineprocess block (1903).

From decision block (1922), if the plant output is at the upper limit ofthe transition zone of operation, then control flows to (1923), theTransition Control subroutine. An exemplary embodiment of thissubroutine is illustrated in the flowchart of FIG. 20. Control thenreturns to the end subroutine process block (1903).

From decision block (1904), if all of the plant's GTs are not operating,then process flow proceeds to decision block (1910). At this juncture,the controller determines whether or not either the ST or HRSG isoperating at a lower limit. For the HRSG and ST, these limits would bedetermined by the plant engineers who would specify the optimum point tobegin shutdown of a GT. If neither of these limits is reached, thencontrol flows to process block (1911), where the fuel flow to the HRSGsis decreased. Control now returns to the end subroutine process block(1903).

From decision block (1910), if the GT or HRSG is at a lower limit ofoperation, then process control proceeds to decision block (1912). Ifthe plant output is greater than the upper limit of the transition zoneof operation, control flows to process block (1914) where load isdecreased on one or more GTs. Control now returns to the end subroutineprocess block (1903).

From decision block (1912), if the plant output is at the upper limit ofthe transition zone of operation, then control flows to (1913), theTransition Control subroutine. An exemplary embodiment of thissubroutine is illustrated in the flowchart of FIG. 20. Control thenreturns to the end subroutine process block (1903).

Transition Zone Operation

Referencing FIG. 20, the Transition Control subroutine begins at (2001)and proceeds to decision block (2002). If a load increase is desired,the process control proceeds to decision block (2010).

If the plant is at the lower limit of the transition zone of operation,process control proceeds to (2011) and an additional GT is started andbrought on line. Control then returns to process block (2012). At thispoint, plant load is modulated by prescribed, programmed outputs for theGTs and STs for a particular transition zone output. Control now returnsto the end subroutine process block (2030).

If a load decrease is required, the process control proceeds to decisionblock (2020).

If the plant is at the lower limit of the transition zone of operation,process control proceeds to (2021) and a GT is taken off line andshutdown. Control now returns to the end subroutine process block(2030).

If the plant is not at the lower limit of the transition zone ofoperation, process control proceeds to (2022) where load is modulated byprescribed, programmed outputs for the GTs and STs for a particulartransition zone output. Control now returns to the end subroutineprocess block (2030).

Summary

The preceding method of controlling the HRSGs and power plant hasillustrated how the teachings of the present invention can beadvantageously applied to power plant operations. It should be notedthat the exemplary system control flowcharts of FIGS. 16-20 may beaugmented or trimmed of steps with no loss in generality or scope ofteachings in regards to the present invention.

The gist of the present invention is that while a large number ofcontrol schemes may be employed to achieve overall cost andenvironmental savings, the basic use of single (or near single) pressureHRSGs in conjunction with supplemental firing can improve the overalleconomics and environmental costs of existing plant technologies.Furthermore, the novel disclosed method of maximizing power plantoperation over a wide range of load while still maintaining the GTs atfull load operation (as contrasted with the prior art) makes thedisclosed control technique a significant improvement in power plantcontrol system engineering.

Preferred System Context of the Present Invention

The numerous innovative teachings of the present application will bedescribed with particular reference to the presently preferredembodiment, wherein these innovative teachings are advantageouslyapplied to the particular problems of a HIGH POWER DENSITY COMBINEDCYCLE POWER PLANT. However, it should be understood that this embodimentis only one example of the many advantageous uses of the innovativeteachings herein. In general, statements made in the specification ofthe present application do not necessarily limit any of the variousclaimed inventions. Moreover, some statements may apply to someinventive features but not to others.

Retrofit Applications

Today, many nuclear, coal, and oil-fired power plants are still inoperation. With increasing pressure to be efficient in a competitiveelectrical marketplace, along with environmental concerns for theproduction of greenhouse gases and other pollutants, the retrofit ofthese existing steam turbine power plants to combined cycle power plantsbecomes more and more likely. However, conventional combined cycle powerplants produce steam at three pressure levels, while the existing steamturbines at conventional steam power plants are designed for utilizingonly HP steam.

In GE informative document GER-3582E (1996), entitled “Steam Turbinesfor STAG™ Combined Cycle Power Systems”, by M. Boss, the authordescribes a basic difference between a ST in a conventional steam powerplant versus a ST in a conventional combined cycle power application:

“Mass flow at the exhaust of a combined cycle unit in a three-pressuresystem can be as much as 30% greater than the throttle flow. This is indirect contrast to most units with fired boilers, where exhaust flow isabout 25% to 30% less than the throttle mass flow, because ofextractions from the turbine for multiple stages of feedwater heating.”

This stated phenomenon greatly complicates the retrofit of conventionalsteam power plants to conventional combined cycle power plants in theprior art. Since conventional power plants accept steam at the inletonly, at HP pressure, they are not designed to accept the IP and LPsteam produced from conventional combined cycle HRSGs. In order to beeffective, it has already been discussed that conventional combinedcycle power plants in the prior art have a ST to GT power ratio ofapproximately 0.5:1. Therefore, to retrofit a 400 MW conventional steampower plant to a conventional combined cycle would require 800 MW of GTcapacity, bringing the total plant capacity to 1200 MW. The existinginfrastructure, fuel lines, available real estate, and most importantly,high voltage power lines, may not be of sufficient size or rating toallow such an uprate (a 200% increase).

In addition, to obtain the high levels of efficiency for the combinedcycle from the prior art, the ST would need to be modified to accept IPand LP steam, and would need to have its entire steam path (internalcomponents including rotating and stationary blades) modified, as theratio of exhaust steam to throttle steam would change from 0.75 in theconventional steam power plant to 1.30 in the conventional combinedcycle power plant. This is a change of 1.3/0.75 or 1.73. This is a majorchange to the steam path of the ST that is very costly and perhaps evenrestrictive, as the present turbine casings may not be usable in aredesign. To further complicate matters, much of the existing equipmentat the existing steam power plant (condensers, pumps, piping, etc.)would no longer be correct for the conventional combined cycleconfiguration. Items such as feedwater heaters are not even used in theprior art combined cycle.

Many of the preferred embodiments of the present invention, however, arean ideal solution to the retrofit option of conventional steam powerplants to combined cycle technology. Since several of the preferredembodiments of the present invention specify the production of primarilyHP steam, this is an ideal option for this retrofit. The currentcombined cycle technology produces steam at up to 1800 psig, while atypical utility standard for steam power plants is 2400 psig, onepreferred inlet pressure for several of the preferred embodiments. Inaddition, since the present invention can utilize a higher ST to GToutput ratio (for example, approximately 1.2:1.0), only 330 MW of GTcapacity is required to retrofit a 400 MW conventional steam power plantto become a clean, efficient combined cycle power plant as described byseveral of the preferred embodiments of the present invention. Also,much of the conventional steam power plant equipment, including the ST,feedwater heaters, condenser, pumps, and other auxiliaries could be usedwith little or no modification.

Retrofit Comparison—Preferred Embodiment to Prior Art

In U.S. Pat. Nos. 5,375,410 and 5,442,908 Briesch and Costanzorespectively propose a hybrid style power plant suitable for use inretrofit applications, but still utilize a three pressure level HRSG.However, supplemental firing is not utilized, and neither is cooling ofthe HRSG exhaust gases by feedwater. Such retrofit power plants operateas a conventional combined cycle when boiler fuel is not used. Incontrast, the preferred embodiments of the present invention utilizeboiler fuel and/or HRSG supplemental firing to determine the bestbalance between fuel types, fuel economics, part load requirements,and/or plant emission levels.

An example for comparison of retrofits for existing steam plants isillustrated in FIG. 44. In this example, an existing steam plantdesigned for standard steam conditions of 2400 psig inlet pressure witha single reheat and inlet/reheat temperatures of 1050° F. is toavailable for retrofit. These steam conditions would normally beassociated with a fossil-fueled power plant, such as coal or oil fired.Although the plant's steam turbine is in good condition, the plant maybe having difficulty with environmental permits, facing expensive boilerrepairs, or be concerned with economic factors in a deregulated powergeneration market. Any one or combination of these factors could beincentive for the plant owners to consider a retrofit of the existingpower plant to the cleaner and more efficient combined cycle technology.

The conventional steam power plant is rated at 400 MW and has a heatrate of 7620 BTU/kWh. If fuel is expensive, it will be advantageous toupgrade this facility to combined cycle technology. However, this plantmay (partly due to its lower heat rate) have a low appraised value. Forthis example it is assumed that this plant has a value of US$50 million,which equates to only US$125 per kW. With low fuel costs, retrofit maynot be economical.

To design an economical retrofit, it is necessary to select the bestequipment combinations that maximize the ST efficiency and capability.For a large ST such as the one in this example, its construction wouldbe similar to that shown in FIG. 51. As can be seen from thisillustration, the rotating and stationary blades in the HP/IP casing tothe left of the figure are much smaller than those in the LP casings tothe right of the figure. Although it is possible to change the bladingin the LP casing, it usually requires a change in the LP casing, whichaffects the foundation, support structure, and condensers. Thefoundation, support structure, and condensers associated with the LPcasings are large heavy components that are difficult and expensive tomodify. Therefore, it is desirable to utilize the ST LP casings withlittle or no modifications, and make steam path changes only to theHP/IP section.

To maximize the existing ST LP section, it is desirable to match itsexhaust flow in the new combined cycle application to that of the formersteam plant, approximately 1,587,000 lb/hr. Utilizing the industrystandard General Electric (GE) Model PG7241(FA) Gas Turbine as the GTengine for this uprate, the total steam production from a 3 pressurelevel HRSG used with this GT would only be 528,000 lb/hr. Therefore, 3GTs of this model would be required in the prior art to effect thisretrofit. This new combined cycle plant from the prior art would berated at approximately 800 MW with a heat rate of 6040 BTU/kWh. However,due to the substantially reduced flows in the HP/IP section of theexisting ST, the blading in these sections would need to be modified.Also, due to the lower volumetric flows, the ST inlet pressure would bederated to 1800 psig. The rating of the modified ST would beapproximately 300 MW. Note that since the combined cycle from the priorart doesn't utilize feedwater heaters, these devices would be isolatedfrom service. Total plant modifications, including those to the HP/IPsection of the ST would be extensive and costly, and US$10 million hasbeen allotted to account for these ST modifications.

Utilizing the technology described by the preferred embodiment on thepresent invention, there are at least two options for this retrofit,demonstrating the flexibility that is offered by the invention. Thefirst option utilizes only one industry standard General Electric (GE)Model PG7241(FA) Gas Turbine and HRSG. This option requires a great dealof supplemental firing, but also produces a great deal of steam. Withmatched exhaust flow to the conventional steam plant, the flows to theinlet of the ST are approximately 93% of the conventional steam plantdesign. Therefore, this ST can be used without modification, with only a7% reduction in inlet pressure at rated conditions. In addition, thisdesign will make use of the existing feedwater heaters. The rating ofthe modified ST would be approximately 375 MW, with a total combinedcycle plant heat rate of 6235 BTU/kWh.

The second option utilizes two industry standard General Electric (GE)Model PG7241(FA) Gas Turbines and HRSGs. With this option, the exhaustflow of the ST exceeds its former design by about 15%. Therefore, theexhaust pressure will climb by about this same amount and overallefficiency will be decreased. In this option of the preferred embodimenton the present invention, inlet steam flows are 87% of the steam plantdesign value, therefore, the ST can be utilized without modification,but with a reduction in inlet pressure at design conditions. This designwill also make use of the existing feedwater heaters. The rating of themodified ST for this second option would be approximately 395 MW, with atotal combined cycle plant heat rate of 6060 BTU/kWh.

FIG. 44 tabulates the data for the various retrofit options. Asdescribed previously, the ultimate determining factor for the retrofitwill be the economic evaluation. If either fuel costs or the plantutilization factor are extremely low, the retrofit may not be warranted.Higher fuel costs may dictate a more efficient plant, but still one withreasonable cost. Limitations on fuel delivery, power line capacity, orreal estate may place restrictions on the power output or the amount ofequipment. Ultimately, the preferred embodiment of the present inventionoffers more options, better utilization of the existing ST, lessinfrastructure change, and lower cost than the retrofit combined cyclepower plant from the prior art.

Combined Cycle Power Plants

The present invention is particularly amenable to application incombined cycle power plants, where the current trend is toward gas-firedcombined cycle turbine systems. The features of the present inventionare attractive in this configuration particularly because of the reducedhardware, space, and capital costs using the teachings of the presentinvention. For example, it is entirely feasible using the teachings ofthe present invention to design a high power density combined cyclepower plant having an initial capital cost which is 25% lower than anequivalent prior art combined cycle configuration.

For example a US$340 million (reference FIG. 25) conventional combinedcycle power plant from the prior art which can be constructed throughthe methods described by the preferred embodiment of the presentinvention, could be built for US$240 million in (reference FIG. 27)capital costs. Initial savings are US$100 million dollars. These savingsequate to US$10 million annually in financed capital costs assuming an8% interest rate amortized over 20 years. Assuming fuel costs for a 725MW plant from the prior art of US$93.4 million per year (reference FIG.25), the annual savings of US$10 million in capital costs equates to10.7% of the total annual fuel costs for the plant. This means that thepresent invention can be up to 10.7% less fuel efficient than currentcombined cycle configurations and still be more economically viable.Obviously, the goal of the present invention is to be as fuel efficientand as environmentally efficient as possible. Thus, the cost savingsover the life of the plant can be significant.

In many new power plant constructions or especially in situations wherethe power plant is a retrofit or upgrade to an existing installation,the amount of real estate available to construct the new plant is fixed.Thus, the present invention capability of providing an equivalent amountof power output with less plant real estate becomes very attractive,especially when overall plant efficiencies can be maintained at or abovecurrent levels.

Furthermore, the ability of the present invention to operate efficientlyover a wide range of part loads is a drastic improvement over the priorart, both from a fuel efficiency standard as well as an exhaustemissions standpoint. Finally, the ability of the present invention whentargeted toward this application to reduce the overall heat rejection ofa high capacity power plant is extremely attractive in light of thenegative impact that this waste heat has on the environment, especiallyconsidering recent scientific studies concerning global warming and thelike.

Energy Transport Fluids

As will be well known by one skilled in the art, while the preferredembodiments have made use of energy transport fluids (ETF) comprisingprimarily air in the topping cycle and steam and/or hot water in thebottoming cycle, the present invention is amenable to application with awide variety of other energy transport fluids such as ammonia,chlorinated fluorocarbons, oil, etc.

These are just a few of the exemplary energy transport fluids that willwork in some context with the present invention, and any mention of“energy transport fluid” or “ETF” should be given its broadest meaningwhen interpreting the intended applications in which the teachings ofthe present invention are germane.

Combustible Fuel and/or Fuel/Heat Sources p As will be well known by oneskilled in the art, while the preferred embodiments have made use ofcombustible fuel (CF/CFT/CFB) comprising primarily natural gas, thepresent invention is amenable to application with a wide variety ofother combustible fuels such as hydrocarbon based fuels, fossil fuels,fuel oil, diesel fuel, and jet fuel. Of course, combinations of singlecombustible fuels may be either mixed and fired or fired separately togenerate a hybrid combustible fuel system that would also be within theanticipated scope of the present invention. These are just a few of theexemplary combustible fuels that will work in some context with thepresent invention, and any mention of “combustible fuel” should be givenits broadest meaning when interpreting the intended applications inwhich the teachings of the present invention are germane.

Similarly, any mention of the term “fuel/heat source (FHS)” whilespecifically including heat generated from the combustion of naturalgas, may also include heat generated from any combustible fuel(CF/CFT/CFB) as defined above, but also may comprise in whole or in partheat derived from a geothermal source, nuclear reactor, nuclear fission,indirect combustion and/or other source of energy.

GT Engine Availability

With the onset of electrical deregulation, there has been a flurry ofactivity by power plant developers to be the first to the marketplacewith new capacity. The business strategy for these developers is thatafter enough power plants have been constructed in a particular region,the banks and other financial institutions will be reluctant to financeadditional power plants in that region. Therefore, the general consensusseems to be that he who builds his plant first, wins the economic race.

This rush to the marketplace has had an effect on the GT manufacturers.At the current time (2nd quarter of 1999) there is approximately a 3year wait for a GE frame 7 GT. In recent years, the lead time for one ofthese GTs was less than 10 months. This is also noted in POWER MAGAZINE,(ISSN 0032-5929, March/April 1999, page 13):

“Gas turbines, which have sold at a modest clip for the past few years,suddenly are selling like stocks with a “dot-com” address, as regulatedutilities and independent power producers (IPPs) rush to developcapacity throughout North America. Some companies are placing orders fordozens of turbines, locking up production slots of the majormanufacturers for years to come.”

This spike in demand for GTs has not only increased the selling price ofmost GTs by a considerable margin, but has made it difficult to evenpurchase some models of GTs without a 2-4 year wait for delivery.Therefore, the preferred embodiment of the present invention serves tocircumvent this problem by producing more power in the ST. Thisalleviates the need for such large amounts of GT capacity, and in someexemplary preferred embodiments, twice the capacity can be attainedwhile utilizing the same GTs that would have been used in a combinedcycle from the prior art.

Westinghouse Model 501G GT Engine

The Westinghouse model 501G gas turbine engine is the next step intechnology from the “F” class engines (includes GE frame 7FA andWestinghouse 501F). The “G” technology engines have higher pressureratios, more sophisticated turbine blade materials, and a firingtemperature of 2600° F. To avoid serious thermal distortion or otherdamage due to high temperature in the combustor/turbine section of theseGTs, it is necessary to provide steam into the gas turbine for coolingpurposes. Thus, in this new technology, the GT is dependent upon thesteam cycle for proper operation. This equipment arrangement hasprovided for higher overall combined cycle efficiencies at fall loadpower, however, there are numerous drawbacks to this technology. Some ofthese drawbacks are listed below:

1. This technology is not yet proven.

2. The cycle does not offer a great deal of flexibility, as supplementalfiring is limited to less than 10% power augmentation. Additionally,this supplemental firing lowers the overall plant efficiency.

3. With the higher combustion temperatures, NOX is more readily formed,and anticipated NOX levels are 42 PPM on natural gas versus only 9 PPMfor a GE frame 7FA GT.

4. With the integral steam cooling of the model 501 G combustionsection, comes the requirement for ultra pure steam. Since the steamcooling passages in the GT components are small, deposits and build-upthat can result from steam impurities are not tolerable. Therefore,special condensate polishing systems are required to produce this highlypure steam.

5. An examination of a heat balance for a model 50IG indicates that someof this cooling steam is consumed in the GT (probably traveling into theturbine section).

For a 2X1 501G combined cycle plant this appears to be 35,000 to 45,000lb/hr of steam. This increases the make-up water requirements, increasesthe duty on the condensate polishing systems, and may be subject toincrease with time as the small passages which leak this steam increasein size due to thermal distortion, erosion, or other factors, thusdegrading the efficiency.

6. Most combined cycles operate with a sliding pressure on the steamcycle to improve efficiency. However, the cooling steam, which emanatesfrom the IP boiler on the HRSG, must be maintained at nearly constantpressure for adequate cooling. This will have detrimental effects onefficiency at part load conditions compared to even conventionalcombined cycle power plants in the prior art.

7. Due to the higher pressure ratio, the model 50IG requires a fuel gaspressure of 600 to 650 psig, versus 350 to 370 psig for a GE frame 7FA.Many pipeline companies will not guarantee pressures to satisfy themodel 501G requirements, so fuel gas compressors are needed in manyapplications where they would not be required for the “F” technologyengines.

8. These GTs require more than 3 hours to reach full load, versus lessthan 30 minutes for “F” technology engines. This limits their use inproviding peak power demands.

As can be seen from this list of drawbacks, the newer technology engines(including the proposed GE “H” technology engines), have a host of newschemes to enhance combined cycle efficiency by a few percent, butrequire a vast amount of restrictive, expensive, and complicatedtechnology to achieve these relatively small incremental increases inefficiency. Although the preferred embodiment of the present inventioncan be used with some of these more advanced engines like the model 501G(however, some changes would be required for cooling steam), many of theexemplary preferred embodiments have focused on the GE frame 7FA andother commercial GT systems due to their proven history, simplicity, lowemissions, and improved efficiency when packaged with the cycledescribed by the preferred embodiment of the present invention.

Combined Cycle Comparison: “G”/“H” GT Technology vs. “F” Technology

In light of the impending deregulation of the electric power generationmarketplace and the subsequent competitive economic environment thatthis deregulation will spawn, the electric power generation industry hasmigrated towards a more sophisticated and complicated means of powergeneration. Specifically, “G” and “H” GT technologies have become thepreferred GT based combined cycles for many proposed combined cyclepower plant installations.

However, the use of this technology will not be without its drawbacks,both economic and environmental. Specifically, the “G” and “H” GTtechnologies provide less operational flexibility than their previous“F” technology counterparts. These newer technologies require amandatory integration of the GT and ST cycles, as the newer GTs requiresteam cooling of internal GT components. Without this ultrapure,precisely metered cooling steam, these GTs will not operate. Therefore,as the combined cycle plant load changes, the steam cycle will not beable to respond as well as even in the prior art, as cooling steamrequirements will dictate the conditions of some steam that is produced.

For control, these new technologies still focus plant operation onmodulation of GT operation to meet plant load requirements, just as inthe prior art. However, due to the nature of their integrated cycles,little or no supplemental firing will be allowed using this technology.This characteristic, when coupled with the plant requirements duringpart load conditions, results in substantially decreased part load heatrates even as compared to older “F” technology plants where there is nodirect coupling between the GT cooling and ST operation. Thus, thesenewer technology GTs are generally designed to be base loaded powerplants. This is in contrast to much of the new plant demand load, whichvaries on a daily and seasonal basis.

Additionally, these newer technology plants have higher firingtemperatures, resulting in the need for more exotic materials in theirconstruction. These higher temperatures therefore lead notably to highermaintenance costs, and also higher NOX emission levels.

Additionally, these newer GTs to achieve the higher efficiencies,utilize higher engine pressure ratios. This results in the need forhigher natural gas inlet pressures, requiring the addition of fuel gascompressors in many situations. These fuel gas compressors consume agood deal of power, and serve to lower efficiency, increase cost, andreduce reliability of combined cycle power plants.

In light of the constraints on operational flexibility, part loadefficiency, increased NOX levels, potential fuel gas compressionrequirements, along with the fact that these “G” and “H” technologymachines have not been proven in even short term operation, the presentinvention has focused on the use of older GT technologies such as the“F” technology. In doing so, the present invention permits decoupling ofthe gas turbine and steam turbine cycles while simultaneously allowingthe GTs to operate at peak fuel and emission efficiencies. The presentinvention using “F” technology provides a power plant that drasticallyimproves part load efficiency, improves plant flexibility, lowersemissions, and drastically lowers overall installed plant cost.

With a heat rate of 6006 BTU/kWh, for a preferred embodiment of thepresent invention, versus 5830 BTU/kWh for Westinghouse “G” technologyand 5690 BTU/kWh for GE “H” technology, this represents only a 3% and a5.5% increase in efficiency at rated load for these more sophisticated(yet operationally limited) combined cycle plants from the prior art.Given the lower part load efficiencies, added maintenance costs,increased capital costs, and lack of operational flexibility, it isunlikely that the “G” and “H” technologies (even with theirincrementally higher full load efficiencies) will provide the economicbenefits available via use of the teachings of the present invention asapplied to combined cycle power plants.

Although the teachings of the preferred embodiments of the presentinvention focus on “F” technology GTs, they may be applied to the “G”and “H” technologies, but only with careful guidance by the GTmanufacturers. Note, however, that the teachings of the presentinvention do not specifically limit application to any particular GT orGT manufacturer, but are valid throughout the range of commerciallyavailable GTs, as are known by one skilled in the art.

Preferred Embodiment Plant Design Method

Since the preferred embodiment consists of a more flexible design for acombined cycle, it offers high efficiency (both at full and part load),and has significant cost advantages associated its high power densitydesign. This method for selecting the optimum power plant for operationand financing is described in subsequent sections below.

Selection

Referring to the exemplary flowchart of FIG. 47, the process begins atthe start block (4701), and continues to decision block (4702), where itis determined whether to investigate new construction or the retrofit ofan existing plant. If the plant will be new construction, processcontrol flows to decision block (4704). If the plant is to be a hybridfuel design, process control proceeds to the Hybrid Fuel DesignSubroutine (4705). Otherwise, process control continues to (4706), wherethe plant developer, using information in (4707) and other informationabout his proposed power plant site such as transmission line capacity,real estate availability, and the commercial value of electricity, willselect a desired combined cycle plant rating (CCR).

Knowing the CCR, the plant developer will proceed to (4708) and,utilizing the input data from (4709), select the GTs for a preferredembodiment combined cycle from a list of selections, such as thatillustrated in FIG. 29 (note that FIG. 29 is only a partial exemplarylist for demonstration purposes). With the GTs selected, the total gasturbine power output, GTP, can be determined. Proceeding to (4710), thesteam turbine power, STP, can be determined as CCR-GTP.

Knowing GTP and STP, process control flows to (4711) where the STP/GTPratio is calculated. Process control now proceeds to (4712) where thedesired efficiency and steam conditions are determined based upon acharacteristic curve similar to that illustrated in FIG. 30. Processcontrol now proceeds to (4801) for an economic evaluation of theselected combined cycle.

Economic Evaluation

Referring to FIG. 48, the economic evaluation begins at (4801) andproceeds to block (4802), where inputs for the load profile, fuel types,fuel cost, and other contributing factors listed in (4803) are used todetermine fuel costs and average annual specific fuel cost in $/kWh.

The process continues to (4804), where inputs for the equipment cost,installation, financing, and other contributing factors listed in (4805)are used to determine capital costs and average annual capital cost in$/kWh.

Process flow continues to (4806), where inputs for inventory cost,maintenance, tools, and other contributing factors listed in (4807) areused to determine maintenance costs and the average annual maintenancecost in $/kWh.

The process flows to (4808), where inputs for personnel cost, taxes,insurance, and other contributing factors listed in (4809) are used todetermine miscellaneous costs and the average annual miscellaneous costsin $/kWh.

Utilizing the data for fuel, capital, maintenance, and miscellaneouscosts, along with the factors listed in (4811), a complete “Economic ProForma” is determined for the proposed combined cycle plant from thepreferred embodiment of the present invention.

The process continues to decision block (4812) to determine if theoption selected is acceptable. If so, process flows to (4813) where thisoption is compared to other acceptable options. Process control proceedsto decision block (4814). If the option calculated is proven to beadvantageous over other acceptable options, it becomes the preferredoption and is saved as such in (4815). Process control continues todecision block (4816). If the new option is not preferred, processcontrol continues to decision block (4816), bypassing (4815).

From decision block (4816), if more options are desired, process controlreturns to the Design/Financing process (4701) in FIG. 47. Otherwise,process flows to (4817) where the preferred option is selected as thebusiness plan for the combined cycle project and process flow then endsat (4818).

Retrofit Plants

Referring to FIG. 49, the Plant Retrofit Process begins at (4901) andproceeds to decision block (4902), where it is determined whether theretrofit is for a hybrid fuel plant or not. If the plant is to be ahybrid design, process flows to the Hybrid Fuel Design Subroutine(4903). After return from this subroutine, the process flows to (4904)to determine the plant economics (see FIG. 48).

If the plant is not a hybrid design, control proceeds to decision block(4905). At this juncture, it must be determined if the existing ST willbe modified (new steam path) or used “as is”. If is to be modified, theprocess goes to (4906) where the new ST rating is determined utilizinginputs from (4907). From here the process returns to (4908). Fromdecision block (4905), if the ST is to be used “as is”, then processcontrol proceeds to (4908).

Using inputs from (4909), the ST rating in the proposed combined cycleis determined and the process continues to (4910). With inputs for fuel,capital, and other contributing factors listed in (4911), a ST/GT powerratio is selected. Proceeding to (4912), utilizing data similar to thatillustrated in FIG. 29, the GTs can be selected. The process nowcontinues to (4801) which is the determination of plant economics (seeFIG. 48).

Hybrid Fuel Plants

Hybrid fuel plants can utilize a number of combustible fuels to provideenergy, as well as nuclear, geothermal, or other heat sources. Byintegrating the combined cycle described by the preferred embodiment ofthe present invention along with the hybrid fuel cycle, improved overallefficiencies and economics are possible.

Referring to FIG. 50, the Hybrid Fuel Design Procedure begins at (5001).Control flows to decision block (5002) where the process decides whetherthe hybrid will use combustible fuel or a heat source like nuclear orgeothermal. If combustible fuel is to be used, process flows to (5005)where the GTs are selected for the hybrid plant based upon relative costof fuels, ST size, desired plant rating, and other contributing factorsas indicated in (5006). From here the subroutine returns to the point ofinvocation.

From decision block (5002), if a heat source like nuclear or geothermalis to be used, the process flows to (5003) where the GTs are selectedfor the hybrid plant based upon relative cost of fuels, ST size, desiredplant rating, and other contributing factors as indicated in (5004).From here the subroutine returns to the point of invocation.

Options

General

As noted in previous discussion, one of the prime advantages of thepreferred embodiment of the present invention is flexibility. This isnot only apparent in the selection of the combined cycle plant rating,but also in its ability to manifest other power solutions such as theretrofit of existing steam plants or the integration of cycles withhybrid fuels. Following is a list of other options that can beeffectively utilized in the preferred embodiment of the presentinvention.

Equipment Arrangement

In U.S. Pat. No. 5,649,416, James H. Moore describes various equipmentarrangements which include GTs and STs coupled together driving a commongenerator. Although the arrangements in FIG. 26 illustrate the GTs andST each with its own respective generator, there is no reason to insistthat this arrangement be required. The teachings of the preferredembodiment are for a new system and method, and the equipmentarrangement could very well be as described by Moore in his patent, oranother arrangement if so desired. Thus, any combination of single-shaftsystem configurations are anticipated by the present invention.

Other Topping/Bottoming Cycles

The present invention has been discussed primarily with respect to theuse of conventional Brayton/Rankine cycles for the combined cycleapplication discussed herein. However, it should be noted that theteachings of the present invention are equally applicable to the use ofother cycles. While there is no practical limit as to what other cyclesmay be utilized within the context of the present invention, it isspecifically anticipated that the GE Kalina cycle (a bottoming cycle)may be particularly amenable to use in conjunction with the presentinvention.

Thus, for the purposes of this document, the terms “topping cycle” and“bottoming cycle” should be given their broadest possible meaningsconsistent with the use of Brayton, Rankine, Kalina, and other cyclesavailable to one skilled in the art. Additionally, it should be notedthat the present invention specifically anticipates the use of multiplecycles within a given combined cycle application.

Small Steam Turbine Driven BFP

For illustrative purposes, the boiler feed pumps (BFP) referenced inthis disclosure are assumed to be driven by electric motors. However, inlarger steam power plants, these pumps are frequently driven by smallsteam turbines, referred to as boiler feed pump turbines (BFPT). TheBPFTs have several advantages over motors, but the primary advantagesare load response and a reduction in exhaust end blade loading.

Since these BFPTs utilize low pressure steam at their inlets (typicallyless than 200 psia), they typically consume a fair amount of steam. Thissteam used by the BPFTs equates to a reduction of steam to the LPsection of the main ST. This reduces the loading on the last stageblades and can often increase the efficiency of the cycle. AdvancedSteam Conditions

In U.S. Pat. No. 5,628,183, Rice discusses development work beingconducted in Europe on higher steam temperatures and pressures, and inthe United States through the Department of Energy (DOE) and theElectric Power Research Institute (EPRI). These include work by SolarTurbines on a pilot project designed for higher cycle efficiencies byutilizing 1500° F. ST inlet steam temperatures. Although not proven inlong term reliable service, as these higher steam pressures and/ortemperatures prove reliable, this technology will be easily implementedinto the preferred embodiment of the present invention.

Advances in GT Technology

Gas Turbine technology continues to improve with advances like moreefficient compressors, new metallurgy, higher firing temperatures,higher pressure ratios, and other efficiency enhancements. As these GTadvances become available, they should be able to be integrated into thecycle herein described by the preferred embodiment of the presentinvention.

Non-Corroding LP HRSG Section

The detrimental effects of GT exhaust gas condensation and its abilityto corrode tubes and fins in the HRSG LP section has been discussed. Onecommon way to avoid this condensation problem is to provide preheatedfeedwater to the HRSG, such that the feedwater is sufficiently warm tobe above the dew point of the GT exhaust gases and preclude theformation of moisture on the HRSG heat exchange surfaces. This methodhas been illustrated in some of the exemplary preferred embodiments,including FIGS. 35 and 39.

However, another method that can be utilized is the use of anon-corroding material in the HRSG tubes and fins, typically stainlesssteel. This construction method eliminates the need for feedwaterpreheating, and allows for further cooling of the GT exhaust gases, andthus even greater heat recovery of energy from said gases. The drawback,however, is the added cost for the stainless steel material. In manyinstances, this added cost will outweigh the value of the energy saved.But if fuel prices were high, and material costs relatively low, thisoption may be economically viable.

Combined HP/LP Pump

In order to produce the required pressures in the steam cycle, a pump istypically employed to pump the feedwater to the desired pressures. Inseveral of the exemplary preferred embodiments, including FIGS. 9, 15,35 and 39, dual pumps are indicated for LP and HP service. These pumpsmay be multiple as illustrated or may be a single pump. As with manypumps utilized for this service, they consist of a series of impellersthat sequentially pressurize the feedwater. A single pump housing, withextraction ports at the proper “pressure” (impeller) location canprovide an intermediate pressure feedwater, while the remainder of thefeedwater continues to the HP outlet. Other pump arrangements can alsobe devised. The intent of the preferred embodiment of the presentinvention is not to limit the size or style of pump, but to allow theuse of any pump or combination thereof that provides the requiredservice.

Waste Heat Recovery

Throughout the discussion of both prior art combined cycle power plantsand the features of the present invention there has been mention oflosses that occur due to equipment inefficiencies in the overall system.For example, this might take the form of losses in the generator due tonon-ideal (non-zero) resistance in the generator windings. In general,most of the system losses in any combined cycle power plant can beexpressed in terms of waste heat, or heat that is generated but notconverted to mechanical or electrical energy. Generator losses, boilerfeed pump losses, lubrication oil losses, ambient GT heat radiationlosses, and ST heat radiation losses are just a few of these waste heatlosses in a conventional combined cycle application. In conventional(prior art) combined cycle arrangements, these waste heat sources aregenerally assumed to be present and not compensated for, as in theseplant configurations the cost of recovering the heat is not economicaland there is little incentive to use this low energy waste heat in auseful application.

Because of the excess low level heat contained in the GT exhaust gases,the prior art utilized a multi-pressure level HRSG to maximize heatrecovery. Through the use of continuous supplemental firing, the energylevel at the high temperature section of the HRSG equals or exceeds theenergy content at the lower temperature section, introducing the needfor ST extraction steam fed feedwater heaters, common Rankine cycledevices not utilized in conventional combined cycles from the prior art.

With this increased need for low level heat in the preferred embodimentof the present invention, other sources of heat may be utilized.Referring to FIG. 21, these include the gas turbine losses, GTL (2102),steam turbine generator losses, STL (2110), and other miscellaneouslosses. Now low temperature heat such as heat from engine lube oil,generator heat losses, and GT compartment cooling air can all be used topreheat feedwater and displace the extraction steam used in the lowertemperature feedwater heaters. The use of this heat not only improvesthe plant heat rate, but reduces the heat rejection requirements for theplant.

The present invention is somewhat unique in these circumstances becausethese waste heat sources can be used in conjunction with feedwaterheaters (as illustrated in FIG. 15) to add heat to water that issubsequently superheated within the HRSG. This practical utilization offeedwater heaters was not possible with the prior art, as the HRSG wasused to provide this function in the prior art, and feedwater heatingwould provide no advantage in the prior art combined cycleconfigurations. Thus, the judicious use of feedwater heating withsupplemental firing in some embodiments of the present invention nowprovide a method of efficiently recovering what was in the prior artunrecoverable waste heat.

It should be noted that the ability to recover this waste heat in apractical manner can be a significant improvement in overall combinedcycle efficiency. Consider, for example, the case in which 1-2% of thewaste heat generated by the system is recovered and put to good use inthe overall combined cycle. Remembering that a large 1000 MW combinedcycle power plant will expend approximately US$175 million annually forfuel means that even a 1% increase in overall cycle efficiency willequate to large savings in fuel (US$1.75 million annually). If thisimprovement can be sustained over a 20-year life cycle of the powerplant, the total fuel savings would be US$35 million. Thus, waste heatrecovery using the present invention represents a new potential forimproving the overall economic efficiency of combined cycle power plantsthat was not a practical possibility using the prior art.

It should not go unnoticed that the recovery of waste heat represents adirect improvement in overall thermal conversion efficiency in thecombined cycle power plant, resulting in a direct reduction in warmingof the atmosphere. Given the increasing concerns regarding the effect ofglobal warming on our environment, an emphasis on waste heat recovery bypower plant designers should be a concern on par with the reduction ofNOX emissions and other forms of pollution. Since it is estimated thatover 100,000 MW of additional electric power plant capacity will be putonline in the next decade, the concerns regarding the waste heatgenerated by these plants will be worthy of inspection by thoseinterested in preserving environmental resources. Additionally, sinceportions of the waste heat generated by combined cycle power plants isexpelled into the environment, there are significant concerns regardingthe impact of this waste heat on both plant and animal wildlife.

Geothermal Plant Augmentation

The present invention may be amenable in many circumstances whereexisting or proposed geothermal power plants which have a low degree ofefficiency are to be augmented with a gas turbine to either (1)supplement the geothermal energy production to meet the desired loaddemand or (2) replace losses or reduction in geothermal energyproduction for existing geothermal power plants. Since the equipmentproduction for a geothermal installation is relatively fixed, the lossof efficiency or energy production in an existing geothermal power plantmay result in the plant being inefficient to operate. In some cases, thereduction in geothermal energy flow may result in a plant shutdown, asthe amount of power being produced may fall below a critical thresholdfor practical plant operation.

The present invention can be advantageously applied to these scenariosin much the same way it is applied to the recovery of waste heat in aconventional combined cycle power plant. The only difference in thissituation is that the ‘waste heat’ used in the present invention isrecovered from a geothermal source. The result of the use of thisgeothermal heat in conjunction with an optimally fired GT results in apower plant that can have a stable power output (regardless of thequality or stability of the geothermal energy source). Since the presentinvention relies heavily on supplemental firing of the HRSG, thegeothermal energy source can in this application be used via heatexchangers to supplant this supplemental firing to the HRSG and thusdisplace the fuel and/or heat normally supplied for this purpose. Thus,as the geothermal energy source declines in output and/or efficiency,this only results in a corresponding increase in supplemental firingfrom other fuel and/or heat sources. The power plant rated outputremains constant, and can even be increased using the retrofit optionsdiscussed elsewhere within this document.

Cogeneration Applications

As mentioned previously, the present invention is particularlyapplicable to cogeneration and combined heat and power (CHP)applications in which both shaft drive and heat are utilized within asingle environment, such as a commercial or industrial plant. In suchapplications, a certain amount of heat from a combined cycle plant maybe used for space heating, chemical feedstock processing, pulpprocessing, paper drying, cogeneration, and/or other industrialprocesses and the like.

The present invention specifically anticipates that the broadestapplication of the teachings of the present invention will be applicableto all forms of cogeneration and CHP applications. As such, the aboveexamples are illustrative only of the range of applications of thepresent invention. Those skilled in the art will no doubt be able toapply the present invention teachings to a wide variety of otherapplications with no loss of generality.

Performance Comparisons

ST/GT Efficiency Tradeoff

To overcome the part load issues associated with electrical system loadfluctuation, several preferred embodiments of the present inventionutilize the steam turbine (ST) as the prime engine. The ST can reduceload easily by closing inlet valves or modulating inlet pressure to theengine (through a change in the rate of supplemental firing). This hasan attenuated effect on part load efficiency as compared to the dilutionof firing temperature as experienced by the GT. Additionally, the ST canactually be designed for optimum efficiency at a designated part loadpoint, where the gas turbine almost always is most efficient at fullload.

An understanding of the differences between gas turbines and steamturbines defines the advantages that STs have in operationalflexibility. Gas turbines consist of a compressor section thatcompresses inlet air (usually at ambient conditions) to anywhere from 3to 30 times atmospheric pressure. This air must then travel to thecombustion zone where it is heated through the combustion of fuel tobetween 1600° F. and 2600° F. at full load, depending upon the GTdesign. These hot pressurized gases then expand through a turbinesection in the GT to produce the power that not only drives thecompressor, but also drives an electrical generator. Approximately ⅔ ofthe power developed by this turbine section is required to drive the aircompressor, while the remaining ⅓ is available to drive the electricalgenerator. Due to the complexity of design, which includes matching thecompressor, combustion system, and turbine section to work as anintegrated unit, GTs are very structured machines. Manufacturerstypically have a variety of models of GTs. However, they are designedfor a distinct output or rating. To obtain a custom designed GT isneither feasible nor economical.

Steam turbines, in contrast, have very flexible designs. They rely uponthe plant boiler feed pumps to provide pressurized water and the plantboiler to provide the heat to convert that high-pressure water intosteam. Therefore, the ST can easily accommodate a change in powerrequirement at the design stage by simply being configured to pass moresteam flow. This is easily accomplished by using incrementally largerstationary and rotating blades in the ST. Typically, a steam turbinedesign engineer can choose from a family of blades in the high-pressure(HP) sections of the ST that may increment as little as 0.25 inches.Blades in the low-pressure (LP) sections of the turbine usually havehigher increment values. Through this design process two different STsmay, for example, have ratings that vary from 100 to 300 MW, and stillfit in essentially the same casing (from the exterior, these twoturbines would look identical). The key difference would be the bladingon the interior of ST and its flow passing capability.

Additionally, by proper selection of the LP blading, it is possible to“overload,” from an efficiency viewpoint, the last stage blades at fullload. Therefore, at full load, these blades are less efficient than atpart load. Then, when the load is reduced, the efficiency of the LPsection actually increases. This design is preferred for plants thatspend a large portion of their operating life at part load, but need toreach peak load for short durations of seasonal peak system demand. Itis this flexibility, along with low maintenance requirements and provenreliability, that make the ST desirable as the prime engine in acombined cycle power plant.

Several preferred embodiments of the present invention define a systemwhereby the exhaust gases enter into an HRSG as in the prior art.However, these exhaust gases typically contain a great deal of oxygen.In fact, the oxygen content of the air is typically reduced from a valueof 21% in ambient air to a range of about 12% to 15% in a typical GTexhaust at full load. This leaves a great deal of oxygen remaining inthe GT exhaust gases to burn additional fuel. If sufficient quantitiesof fuel are burned, all the steam that would have been produced as lowerpressure steam in the prior art, can be upgraded to HP steam with theproper system modifications as described by the preferred embodiment ofthe present invention. In this manner, the HP steam flow is greatlyincreased, and the ST size relative to the gas turbine(s) (ST/GT powerratio) goes from a nominal 0.5 in the prior art to a value typicallygreater than 1.0. Therefore, rather than being primarily a GT cycle witha ST recovery cycle, the present invention is more like a conventionalsteam plant with additional GT power production and the ducting ofexhaust gases from the GT into the steam power plant's boilers topreheat air and increase boiler efficiency. To maximize efficiency inseveral preferred embodiments of the present invention, a completeintegration of the cycles is required, including utilization of wasteheat, feedwater heating, and implementation of controls to optimize heattransfer.

Comparison of Prior Art to Exemplary Preferred Embodiments

As detailed, the prior art combined cycle technology evolved fromsmaller cogeneration plants. In the state-of-the-art combined cyclepower plant from the prior art, the GT exhausts to an HRSG that istypically either two or three pressure levels. The steam from each ofthese pressure levels is then directed to the ST at the appropriatepoint corresponding to the pressure level of the HRSG section.Supplemental firing is utilized as a means to obtain higher output, butthis done only intermittently to meet peak load, and is accomplishedonly with a reduction in thermal efficiency. Primary load control forthe prior art combined cycle power plant is still achieved by modulatingload on the GT(s). Single pressure level HRSGs can also be employed witha corresponding reduction in thermal efficiency. Higher pressure inletsto the steam turbine typically are not justified as the low volumetricflows to the ST offset any cycle gains from higher pressure by reducedturbine efficiency in the HP section.

With the prior art of combined cycle technology, the plant is primarilya GT based plant that added an HRSG to recover the waste heat. The ST isthen designed to make the best of this recovered heat (which isconverted to steam by the HRSG at multi-pressure levels). The typicalST/GT output ratio for these combined cycle plants is in the 40% to 60%range, with a typical number for a GE S207FA plant being approximately0.57. With the need to utilize HP, IP, and LP steam, the ST hasrelatively low flows in the HP section and higher flows in the LPsection. This reduces HP volumetric efficiency and increases therelative size and cost of the exhaust section(s). Feedwater heating isdone in the HRSG and conventional ST extraction steam fed feedwaterheating is not employed. Preheating of the feedwater from the condensermay be utilized, but the purpose for this process is not to enhanceefficiency but to avoid condensation of water vapor in the exhaustgases. Since water is contained as humidity in the inlet air, and isalso formed as a product of combustion of hydrocarbon fuels, thisincreased concentration of water vapor in the exhaust gases lowers thedew point. Cold feedwater direct from the condenser can causecondensation on the economizer tubes and fins. This condensation hasbeen demonstrated to corrode these fins, lessen the heat exchangeeffectiveness, and cause detrimental effects in HRSG performance. Thusthe use of a feedwater preheater may be utilized in some applications.

In summary, the combined cycle plant from the prior art is primarily aGT based plant with the steam cycle designed as a compromise between thebest cycle efficiency and optimum exhaust gas heat recovery. There areoptions such as supplemental firing to increase plant rating by anominal amount (typically less than 25%), but this additional powercomes with a penalty on plant heat rate. Due to the rigidity of designof the GT, there is little flexibility in the rating or design of thecombined cycle plant from the prior art. In essence, the prior art is arigid power plant design based on the GT engine or set of engines, withan HRSG, and a ST rated nominally at 50% of the GT output. The SToperates in a dependent mode and follows the GT load.

In several preferred embodiments of the present invention, the GTexhausts to a single pressure level HRSG (or primarily single pressurelevel) that is designed for continuous supplemental firing. Thissupplemental firing increases the steam production by a significantamount, and subsequently increases the feedwater flows such thatadditional pressure levels in the HRSG are not required to cool theexhaust gases to optimum temperature (approximately 180° F.). Feedwaterflows that exceed the optimum flow through the HRSG are directed toconventional ST extraction fed feedwater heaters to improve steam cycleefficiency. Due to the flexibility of design, the combined cycledescribed by several of the preferred embodiments of the presentinvention has a ST/GT output ratio that can vary from approximately 0.75to 2.25. Of course, those skilled in the art will recognize that otheroutput ratios are also possible and within the scope of the teachings ofthe present invention. For most load variations on the plant, the GT(s)remain primarily at or near their most efficient load (100%) and thesupplemental firing rate is modulated to change the ST load.

In summary, several of the preferred embodiments of the presentinvention are in essence a large central steam power plant similar tothose known in the prior art of steam power plants, with the boilerreplaced by the HRSGs which continuously bum fuel, just like a boiler ina conventional steam plant. However, GTs have now been added to thecycle which provide oxygen rich (12-15%) hot gases to the boiler (HRSG),increasing its efficiency and allowing for the combustion of additionalfuel. Feedwater heating is accomplished in both the HRSG low temperaturesections and in conventional extraction steam fed feedwater heaters. TheST is larger with more mass flow through the HP and IP sections and lessthrough the LP section (steam extracted for feedwater heating reducesexhaust end flow), increasing volumetric efficiency and decreasingrelative exhaust size. Several preferred embodiments of the presentinvention become combined cycle power plants that are more flexible,have improved full and part load efficiency, and are less expensive toconstruct, operate, and maintain.

Major Equipment Maintenance Costs

Besides fuel and capital costs, another large expense for combined cyclepower plants is the cost for maintenance, and especially maintenance onthe major pieces of equipment such as the GT and the ST. Thesemaintenance costs vary with the equipment model, its complexity, anddegree of service (high temperature or low temperature, steady or cyclicduty, etc.). Typically maintenance costs are examined on a mills/kWhbasis, where a mill is US$0.001 or 0.1 cents U.S. currency. Following isa list of expected maintenance costs for some major pieces of equipment,along with the annual expected maintenance costs based upon a normalized200 MW output at 70% capacity (1,500,000,000 kWh per year):

Maintenance Rate Annual Cost Description (mills/kWh) (US $) 2400 psig ST0.5 750,000 GE Frame 7FA GT 2.5 3,750,000 Westinghouse 501G GT 4.56,750,000

As can be seen from the maintenance numbers, it is much more expensivefrom a maintenance perspective to operate a GT than it is a ST. Inaddition, the advanced technology GT (model 501G) with its higher firingtemperature, single crystalline blades, and steam cooled combustionsection, is also projected to be an expensive piece of equipment from amaintenance perspective.

With the prior art, the GTs produce approximately 67% of the power,while the ST produces 33% (ST/GT output ratio of 0.5:1.0). Since the GTsare modulated to change load, and the ST follows, this ratio is fairlyconstant throughout the load range. Therefore, over a year of operation,in a combined cycle power plant in the prior art, the GT maintenancefactor would be applied to 67% of the kWhs produced, and the STmaintenance factor to 33% of the kWhs produced.

With several of the preferred embodiments of the present invention, theratios are not so simple, for as the total plant load changes, the SToutput is modulated to the greatest extent possible, while the GTs aremaintained at or near full load.

Part Load Efficiency Comparisons

The teachings of the present invention can be best explained incomparing the performance comparisons illustrated graphically in FIGS.6, 15, 33, and 22-28.

FIG. 33 graphically illustrates the part load performance differencebetween two state-of-the-art conventional combined cycle power plantsand two preferred embodiments of the present invention. This graphillustrates that the present invention performance is significantlysuperior to conventional combined cycle power plants at part loadoperation. As can be seen from a comparison of the tabulated data inFIG. 25 and FIG. 27, over the typical operation profile, the exemplarypreferred embodiment of the present invention uses less fuel, costsabout US$100 million less to construct, and has NOX emissions which areless than ⅓ of those in the Westinghouse combined cycle power plant.Thus, the present invention embodiment illustrated in FIG. 26 providesboth significant cost savings and simultaneous savings in environmentalpollution due to reduced NOX emissions. This characteristic is generallya feature of the present invention teachings and is in essence the bestof both worlds—economic efficiency with simultaneous reduction inpollution.

To utilize several of the preferred embodiments of the presentinvention, the HRSG must generally be of a more stout construction tohandle the higher pressures and temperatures than required in the priorart. This can be accomplished in numerous ways. First, the use of awater-wall (vertical tubes filled with feedwater) may be needed to linethe combustion area of the HRSG to protect it from the high combustiontemperatures. As an alternative, the exhaust gases could first be cooledthrough the superheater section (to approximately 800° F.), thenreheated to 1600° F. before continuing though the HRSG. Currently, 1600°F. seems to be the upper temperature limit that manufacturers specifyfor standard HRSG construction. Yet another alternative is the use ofdual grids of duct burners in the HRSG. After the GT exhaust gases areheated to 1600° F., they are allowed to cool through the initialsections of the HRSG, then more fuel (heat) is added through combustionat a point downstream. This adds approximately twice the heat as onegrid burner without exceeding any limiting HRSG temperatures.

Conclusion

The present invention permits a wide variety of applications, but itmust be noted in summary that the use of the present invention in thefield of combined cycle power plants is particularly advantageous inlight of current trends in the power generation industry. While theprior art has in general taught away from the use of supplemental firingof HRSGs as a means of increasing overall plant efficiency, the presentinvention has embraced this concept.

Within the context of an overall improvement in system efficiency, thepresent invention promotes the use of supplemental firing not togenerate more steam as in past power generation applications, but thegeneration of more high quality steam. By this it is meant that byexpending additional fuel in the supplemental firing of the HRSG it ispossible to generate steam which is more energetic and thus capable ofmore efficiently generating power when used in conjunction with asuitably designed bottoming cycle engine.

While the present invention when used in the context of power generationwill require the construction of HRSGs which can sustain highertemperatures than are currently the norm, the materials to accomplishthis are readily available and both steam plants and HRSGs in the priorart have demonstrated at these elevated temperature levels. Furthermore,data in this disclosure indicates that in many circumstances these HRSGswill be smaller than existing units, meaning that construction andmaintenance costs may be comparable to or even lower than existingunits. In addition, the fact that the HRSGs in this application may inmany cases be of the single pressure variety may in some circumstancesprovide some economy in their design and construction.

It must be stressed that the potential energy density of the presentinvention has significant ramifications in regard to the amount ofsupport hardware required to implement a power plant. To achievereasonable overall efficiency, conventional plants make use of a numberof GTs and STs so that when these units are operated at part load thatthe overall system can be operated with reasonable efficiency. This isprimarily because operating the GT at a part load is in general veryinefficient. The present invention circumvents this environmentallydetrimental effect by endeavoring to operate all GTs at their optimalefficiency (both economic and environmental), thus allowing feweroperating GTs to achieve the same overall efficiency and environmentalimpact while consuming fewer support resources.

These efficiency enhancements of the present invention are furthersupplemented by a mechanism whereby the capacity of the plant may betemporarily expanded beyond its normal nominal rating, albeit at a lowerefficiency level. This extension of the plant rating to support higherloads can be a critical factor in the economics of power plantconstruction, because the environmental and logistical hurdles that mustbe overcome to actually construct new power plants are becoming theparamount economic issues barring new plant construction. As such, thepresent invention permits the useful performance of a given power plantconfiguration to be extended beyond that of a conventional power plant,thus permitting the plant rating to be dynamically adapted to meettemporary overload conditions. This capability can have dramaticeconomic cost and environmental savings in that the present inventionpermits the incremental economic and environmental costs to be reducedin the face of a demand for a temporary increase in plant output.

Finally, it must be stressed that while past power plant designs haveendeavored to optimize their operation based on fuel costs alone, thepower plants of the future must incorporate and optimize costs ofcapital, environmental impacts, real estate costs, regulatory costs, andthe ever increasing costs of technology and support machinery. It is theintent of the present invention to address all of these factors inunison and obtain an overall plant design that is a cost effective,power efficient, and environmentally friendly method of generatingpower.

Although a preferred embodiment of the present invention has beenillustrated in the accompanying Drawings and described in the foregoingDetailed Description, it will be understood that the invention is notlimited to the embodiments disclosed, but is capable of numerousrearrangements, modifications, and substitutions without departing fromthe spirit of the invention as set forth and defined by the followingclaims.

What is claimed is:
 1. A combined cycle power plant process comprising:providing a topping cycle fluid (“TCF”) from a topping cycle engine(“TCE”) to a heat recovery device (“HRD”); recovering exhaust hear fromthe TCF in the HRD; substantially continuously supplemental firing theHRD; providing a bottoming cycle fluid (“BCF”) from the HRD to abottoming cycle engine (“BCE”); and modulating the rate of BCF flowthrough at least a portion of the HRD such that an exhaust temperatureof the HRD is maintained at a predetermined temperature range, therebycontrolling heat recovery from the TCF in the HRD.
 2. The combined cyclepower plant process of claim 1 wherein said TCE includes at least onegas turbine (“GT”), wherein said BCE includes at least one steam turbine(“ST”), wherein said HRD includes at least one heat recovery steamgenerator (“HRSG”), and wherein said BCF is steam in said ST andfeedwater in at least a portion of said HRSG.
 3. The combined cyclepower plant process of claim 2 wherein modulating the rate of BCF flowincludes modulating a rate of feedwater flow through at least a portionof said HRSG by diverting feedwater from said HRSG to a parallelfeedwater loop and preheating said parallel feedwater using STextraction steam.
 4. The combined cycle power plant process of clam 2wherein said HRSG produces steam at a single pressure.
 5. The combinedcycle power plant process of claim 2 wherein said HRSG produces highpressure steam predominantly at supercritical pressure at rated combinedcycle plant output.
 6. A combined cycle power plant process comprising:providing a topping cycle fluid (“TCF”) from a topping cycle engine(“TCE”) to a predominantly single pressure level heat recovery device(“HRD”); recovering exhaust heat from the TCF in the HRD; supplementalfiring the HRD; providing a bottoming cycle fluid (“BCF”) from the HRDat a predominantly single pressure level to a bottoming cycle engine(“BCE”); and maintaining a TCF exhaust temperature of said HRD in apredetermined temperature range such that heat recovery in said HRD iscontrolled.
 7. The combined cycle power plant process of claim 6 whereinsaid exhaust temperature of said HRD is maintained in said predeterminedtemperature range by controlling BCE flow through at least a portion ofthe HRD.
 8. The combined cycle power plant process of claim 6 whereinsaid exhaust temperature of said HRD is maintained in said predeterminedtemperature range by controlling supplemental firing of said HRD.
 9. Thecombined cycle power plant process of claim 8 wherein the supplementalfiring is substantially continuous.
 10. The combined cycle power plantprocess of claim 8 wherein total supplemental firing input at combinedcycle plant rated capacity added is at least about 30% of the energyinput to said TCE.
 11. A method of operating a combined cycle powerplant comprising at least one gas turbine (“GT”), at least one steamturbine (“ST”), and at least one heat recovery steam generator (“HRSG”)associated with at least one GT, said method comprising: operating saidGT to produce shaft work and exhaust gas; passing said exhaust gasthrough said HRSG; supplemental firing said HRSG, wherein totalsupplemental firing energy input is at least about 30%of the energyinput to its associated GT; producing steam at said HRSG using heat fromsaid exhaust gas and said supplemental heat; passing said steam to saidST; operating said ST to produce shaft work; converting said steam tofeedwater; passing said feedwater to said HRSG; and maintaining anexhaust temperature of said HRSG in a predetermined temperature rangesuch that heat recovery in said HRSG is controlled.
 12. The method ofclaim 11 further including diverting at least some of said feedwaterfrom said HRSG to a parallel feedwater loop and preheating saidfeedwater in said parallel feedwater loop using ST extraction steam.